Subcontract Report

NREL/SR-500-45103

August 2010

Low Wind Speed Turbine

Developments in Convoloid 

Gearing

Final Technical Report

June 2005 – October 2008

Genesis Partners, LP

Montgomeryville, Pennsylvania
National Renewable Energy Laboratory

1617 Cole Boulevard, Golden, Colorado 80401-3393

303-275-3000 • www.nrel.gov

NREL is a national laboratory of the U.S. Department of Energy

Office of Energy Efficiency and Renewable Energy

Operated by the Alliance for Sustainable Energy, LLC

Contract No. DE-AC36-08-GO28308

Subcontract Report

NREL/SR-500-45103

August 2010

Low Wind Speed Turbine 

Developments in Convoloid 

Gearing

Final Technical Report

June 2005 – October 2008

Genesis Partners, LP 

Montgomeryville, Pennsylvania

NREL Technical Monitor: Scott Schreck

Prepared under Subcontract No. ZAM-3-33200-12
NOTICE

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iii

List of Acronyms

AGMA American Gear Manufacturers Association

B&S Brown & Sharpe 

CMM coordinate measuring machine

COE cost of energy 

CV constant velocity

FCR fixed charge rate 

FEA finite-element analysis

GB gearbox 

HS high speed 

IP intellectual portfolio

LS low speed

LVR load distribution analysis 

LWST low wind speed turbine 

O&M operations and maintenance 

PMM precision measuring machine 

VFD variable frequency drive

WT wind turbine 

WTG wind turbine generator 
iv

Executive Summary

Introduction

This report presents the results of a study conducted by Genesis Partners LP as part of the 

United States Department of Energy Wind Energy Research Program to develop wind 

technology that will enable wind systems to compete in regions having low wind speeds. 

The purpose of the program is to reduce the cost of electricity from large wind systems in 

areas having Class 4 winds to 3 cents per kWh for onshore systems or 5 cents per kWh 

for offshore systems. This work builds upon previous activities under the WindPACT 

project, the Next Generation Turbine project, and Phase I of the Low Wind Speed 

Turbine (LWST) project. This project is concerned with the development of more costeffective gearing for speed increasers for wind turbines. 

Approach

The project statement of work required the refinement in the proposed technical approach 

to achieving the LWST cost of energy (COE) objective, focusing on the development the 

Convoloid gear shape—a novel new design approach for gear teeth. (See Appendix V) 

This new design approach is expected to provide gears that will have increased loadcarrying capacity within the same gear envelope of the existing gear design. 

Results

The project statement of work was to test Convoloid gearing in a 4-square test rig based 

on 108-kW wind turbine gearboxes. For the “slave gearbox,” optimized involute gearing 

was manufactured and assembled into the 108-kW housing. Convoloid gearing of the 

same materials, ratio, bearings, and face widths as the involute design (only change was 

in the tooth form) were assembled into the “test gearbox.” 

After 942 hours of testing, the Convoloid gearbox suffered a major bearing failure 

(spherical roller bearings) on the intermediate shaft. Comprehensive analysis showed that 

reconditioning both boxes would be time consuming and expensive when considered in 

the context of the contract constraints. After consulting with NREL personnel, the 

decision was made to scale down the size and increase the cost efficiency of testing. The 

“3-Hole” Test project then was initiated on May 18, 2006. It included the objectives of 

lower required test cycles, potential “24/7” (24 hours per day, 7 days per week) test-rig 

operation, increased numbers of gear-pair specimens (to enhance the statistical surety of 

the results), greatly increased efficiencies in gearbox change-outs, and multiplicity of test 

results. Test results for the NEG108 Micon testing are found in Section 4 and those of the 

3-hole testing in Section 7.

Key Findings

1. A proven method of adapting the non-involute Convoloid tooth form coordinate 

system to the industry's CNC gear tooth grinders was found using the established 

machine software oriented only to involute forms. This advancement makes the 

economical production of Convoloid forms with the world gear industry's present 

capital asset infrastructure possible and does it with ease. 

 
v

2. The tooth tip relief protocols for Convoloid were developed before the tests and 

applied to the Convoloid gears. The low speed, high torque gears were perfectly 

relieved but the high speed lower torque gears had too much relief due to the high 

tooth stiffness of the Convoloid design. Tip relief protocols were altered 

accordingly to match the test results and have been successfully used since. 

 

3. The Convoloid forms developed micropitting at the entry and exit to the transition 

zone. New protocols have been developed to preclude this occurrence in future 

designs. 

 

4. At the conclusion of the tests at 965 hours of operation with substantial segments 

of the tests run at over 4 times rated loads, the Convoloid micropitting had 

steadied out whereas the involute gears showed a steady migration of the 

micropitting phenomenon. This fact proved to the test observers that Convoloid 

gearing will be far superior to involute gearing of the same size, material, face 

width, and center distance. 

 

5. Using only NREL COE protocols it was determined that (paragraph 9.5 of the 

Final Report)--"The energy capture of the cost neutral Convoloid gear box/blade 

diameter increase typified by alternative #3 using Class 3 winds (5.539 m/s) 

equals the energy capture of the classical involute Baseline design using Class 4 

winds (5.8 m/s). This result exactly matches original NREL LWST Program 

Objectives as directed: to improve energy capture of the wind resources found in 

the Midwestern United States."
vi

Table of Contents

List of Figures ...............................................................................................................xi

List of Tables...............................................................................................................xix

1. Introduction................................................................................................................ 1

1.1. Background .........................................................................................................................1

1.1.1. Low Wind Speed Technology ...................................................................................1

1.1.2. Organization ...............................................................................................................2

1.2. Objectives ...........................................................................................................................2

1.3. Scope ...................................................................................................................................2

2. Methodology—Common to Both Micon 108 and 3-Hole Tests .............................. 3

2.1. Work Plan ...........................................................................................................................3

2.1.1. Procedure ...................................................................................................................3

2.1.2. Gear Design Methodology .........................................................................................4

2.1.3. Test Fixture Design Methodology .............................................................................5

2.2. Testing.................................................................................................................................6

2.2.1. Description .................................................................................................................6

2.2.2. Test Gearbox Component Analysis ...........................................................................7

2.2.3. Testing Methods.........................................................................................................7

2.2.4. Test Protocol ..............................................................................................................7

2.2.5. Definition of Failure ..................................................................................................8

2.3. Cost of Energy Analysis .....................................................................................................8

2.4. Convoloid Gear Inspection Methods ..................................................................................9

2.4.1. Hob Inspection ...........................................................................................................9

2.4.2. Gear Inspection ..........................................................................................................9

3. Project Initiation......................................................................................................... 9

3.1. Initial COE Projection.........................................................................................................9

3.2. Project Task Outline .........................................................................................................11

3.3. Stress Comparisons ...........................................................................................................12

3.3.1. Rating Factors ..........................................................................................................12

3.3.2. Projection of Rating Factors and Stress Levels .......................................................12

3.3.3. Project Operation and Maintenance Costs as a Function of Stress ..........................13

3.4. NEG Micon 108 Gearbox Housing Rebuild .....................................................................13

3.4.1. Housing Rigidity ......................................................................................................14

3.4.2. Bearing Bores...........................................................................................................14

3.5. Involute Gearbox Design ..................................................................................................14

3.5.1. Involute Gearbox Component Manufacture ............................................................14

3.5.2. Involute Gearbox Inspection Procedures .................................................................15

3.6. Optimized Convoloid Design Gearbox .............................................................................15

3.6.1. Convoloid Gearbox Component Manufacture .........................................................15

3.6.2. Convoloid Inspection Methods ................................................................................16

4. Testing—Micon 108 ................................................................................................. 18

4.1. Description of Test ............................................................................................................18

4.1.1. Gearbox and Test Fixture Configuration .................................................................18

4.1.2. Test Fixture ..............................................................................................................18

4.2. Test Preparations ...............................................................................................................21
vii

4.2.1. Test Gearbox Design ................................................................................................21

4.2.2. Manufacturing Sensitivities of Convoloid Gearing and Their Effects on 

Rating Factors ..............................................................................................................23

4.2.3. Gear Manufacturing .................................................................................................25

4.2.4. Test Protocols...........................................................................................................26

4.3. Testing...............................................................................................................................27

4.3.1. Early Test Observations—200% x 200-Hour Test Protocol ....................................27

4.3.2. Test Observations—Extended Testing ....................................................................31

4.3.3. Cessation of Test ......................................................................................................34

4.3.4. Variations During Test .............................................................................................40

4.3.5. Load/Power Summary .............................................................................................41

4.4. Test Results .......................................................................................................................42

4.5. Post-Test Analysis ............................................................................................................42

4.5.1. Housing Stiffness / Bearing Redesign for Micon 108 .............................................42

4.5.2. Convoloid Profile Modification (Tip Relief) Study .................................................43

4.6. Test Conclusion ................................................................................................................45

4.6.1. Test Limitations .......................................................................................................45

4.6.2. Conclusions ....................................................................................................................45

4.6.3. Analysis of Test Results—Micon 108 .....................................................................46

5. Change of Test Program Venue..............................................................................49

5.1. Meeting of NREL and Genesis Partners’ Personnel .........................................................49

5.2. Analysis of Change of Test Program Venue .....................................................................49

6. Project Initiation—3-Hole Test................................................................................49

6.1. Description of Test ............................................................................................................49

6.1.1. 3-Hole Gearbox Design ...........................................................................................49

6.1.2. Test Equipment and Fixture Design ........................................................................53

6.2. Test Preparations ...............................................................................................................57

6.2.1. Test Equipment ........................................................................................................57

6.2.2. Test Protocols...........................................................................................................61

7. Testing—3-Hole Test ...............................................................................................62

7.1. Testing...............................................................................................................................62

7.1.1. Early Test Observations—Micropitting Test ...........................................................62

7.1.2. Test Observations—Extended Testing ....................................................................65

7.2. Variations During Test ......................................................................................................70

7.2.1. Incorrect Torque Application ...................................................................................70

7.2.2. Loss of Lubrication ..................................................................................................71

7.3. Test Results .......................................................................................................................71

7.4. Continued Future Testing .................................................................................................71

8. Detailed Design of Larger Gearboxes Using Convoloid Gear Sets.....................71

8.1. Selection Criteria ..............................................................................................................71

8.2. Extrapolation .....................................................................................................................72

8.3. Comparison of Original 750-kW Gearbox with Convoloid Enhanced Gearbox ..............72

8.4. Analytical Model Cost Savings and Extrapolated Results ...............................................74

9. Cost of Energy Analysis—Summary......................................................................74

9.1. Calculation of Achievable Reductions in Size and Cost of Convoloid Versus 

Involute Wind Turbine Generator ......................................................................................74
viii

9.2. Calculations of Advantages of Retrofitting Convoloid Gearing into Existing 

Involute Wind Turbine Generators ....................................................................................75

9.3. Cost of Energy Trade-Offs ................................................................................................75

9.3.1. Wind Energy Generation—Trade-Off 1: Gearbox Cost Reduction / Tower 

Height Increase (Near Net Cost Neutral) .....................................................................76

9.3.2. Wind Energy Generation—Trade-Off 2: Gearbox Cost Reduction / Blade 

Diameter Increase (Near Net Cost Neutral) .................................................................78

9.3.3. Wind Energy Generation—Trade-Off 3: Gearbox Cost Reduction / Blade 

Diameter Increase (Near Net Cost Neutral)—Reduced Wind Speed ..........................81

9.4. Best Trade-Off Model .......................................................................................................82

9.5. Objectives of NREL LWST Development Program ........................................................82

9.6. Potential Effects of Convoloid Gearing Technology ........................................................82

9.6.1. Performance .............................................................................................................82

9.6.2. Economic Factors.....................................................................................................83

9.6.3. Extended Service Life ..............................................................................................83

Appendix A. Three-Hole Test......................................................................................84

Involute Gear Rating per AGMA 2001 ...................................................................................84

3-Hole Test Instrumentation Data Points .................................................................................91

Overview of 3-Hole Fixture Design and Shakedown ..............................................................92

Micropitting Test Protocol .....................................................................................................103

3-Hole Gearbox Test ..............................................................................................................103

First Start Procedure ........................................................................................................103

Micropitting Test Procedure ............................................................................................104

Extended Comparison Test Protocol .....................................................................................105

3-Hole Gearbox Test ..............................................................................................................105

Extended Comparison Test Procedure .............................................................................106

Start-Up Procedure...........................................................................................................108

Shutdown Procedure ........................................................................................................108

3-Hole Test—Detail of Test ...................................................................................................108

Estimated Test Time—Continuation of 3-Hole Test .............................................................122

Involute ............................................................................................................................123

Convoloid .........................................................................................................................124

Early Test Results ..................................................................................................................124

Failure Modes ..................................................................................................................125

3-Hole Test Life Analysis ......................................................................................................126

Low Wind Speed Turbine Baseline Turbine—Baseline Operating and COE 

Parameters ........................................................................................................................127

Operating Conditions and Parameters ...................................................................................127

Appendix B. Baseline COE Comparison Sheet (2002 $).........................................130

Baseline COE Comparison Sheet (2002 $) ............................................................................130

Appendix C. Detailed Task Outline...........................................................................132

Main Objectives of the Program ............................................................................................132

Appendix D. Gearing Life Analysis—Surface Failure Mode...................................136

Definition of Surface Failure .................................................................................................136

Summary of Types of Common Surface Failure Modes .......................................................136

Pitting (Macropitting) ......................................................................................................136
ix

Micropitting .....................................................................................................................136

Wear .................................................................................................................................137

Scuffing ............................................................................................................................137

Plastic Deformation .........................................................................................................137

Appendix E. Micon 108 Instrumentation Data Points .............................................138

Appendix F. Involute Gear Rating per AGMA 2001.................................................140

Low Speed Gear Set ...............................................................................................................140

High Speed Gear Set ..............................................................................................................147

Appendix G. Comparison of Stress Levels At 200% Rated Load ..........................154

Summary of Gear Bending Strength and Surface Durability Ratings ...................................154

Results of the Stress Analysis ..........................................................................................154

Involute ............................................................................................................................154

Convoloid .........................................................................................................................154

Summary ................................................................................................................................155

Appendix H. Micon 108 Gearing Life Calculation....................................................156

Summary of Gear Surface Durability Life .............................................................................156

Results of the Pitting Life Analysis .......................................................................................156

Involute ............................................................................................................................156

Convoloid .........................................................................................................................156

Appendix I. Component Analysis at 240% Rated Load ..........................................157

Appendix J. Manufacturing Sensitivities of Convoloid Gearing and Their 

Effects on Rating Factors ....................................................................................160

Appendix K. Shakedown Protocol—Micon 108 in Preparation for Executing 

the Test Protocol ..................................................................................................163

Appendix L. Micon 108 Test Protocol ......................................................................164

References ..............................................................................................................................164

Appendix M. Test Protocol—Addendum A: Continuation of Test.........................167

Description .............................................................................................................................167

Appendix N. Description of Start-Up and Shutdown Procedures—Micon 108.....168

Start-Up Procedure.................................................................................................................168

Shutdown Procedure ..............................................................................................................168

Appendix O. Micon 108—Detail of Test ...................................................................169

Test Setup...............................................................................................................................169

Crown Specification .........................................................................................................170

Tooth Twisting .................................................................................................................170

Tip Relief .........................................................................................................................170

Test Observations...................................................................................................................171

200% x 200-Hour Test .....................................................................................................171

Micon 108 Extended Test (240% load) ...........................................................................174

Appendix P. Micon 108 200% x 200-Hour Detail of Test Figures and Photos.......189

Appendix Q. Micon 108 Test Recommendations ....................................................207

Appendix R. Micon 108 200% x 200-Hour Test Vibration Analysis........................209

Involute ..................................................................................................................................210

Motor......................................................................................................................................210

Conclusions ............................................................................................................................210
x

Appendix S. Test Observations—Richard L. Meredith, D.C. Energy.....................213

Observations From Genesis ...................................................................................................213

200% Rated Load 235-Hour Test Data ..................................................................................213

Appendix T. Derivation of Gearbox Efficiencies .....................................................214

Overall System Efficiency .....................................................................................................214

Input Power Determination ....................................................................................................214

Circulating Power ..................................................................................................................215

Efficiency ...............................................................................................................................215

Gearing Efficiency .................................................................................................................215

Appendix U. Convoloid Gear Separating Force Qualification................................217

Static Load Test Equipment ...................................................................................................217

Test Results ............................................................................................................................219

Conclusions ............................................................................................................................221

Appendix V. Tutorial on Convoloid Gearing Technology.......................................224

Description .............................................................................................................................224

Trademark ..............................................................................................................................224

Tooth Form ............................................................................................................................224

Architectures ..........................................................................................................................224

Advantages .............................................................................................................................224

Patents ....................................................................................................................................225
xi

List of Figures

Figure 1.1. Organization of project personnel .................................................................... 2

Figure 2.1. Micon 108 back-to-back test stand arrangement .............................................. 4

Figure 2.2. Micon 108 test fixture, involute gearbox on the left and Convoloid gearbox on 

the right ....................................................................................................................... 5

Figure 2.3. Micon 108 test fixture, Convoloid gearbox with torque application hydraulic 

cylinder ....................................................................................................................... 6

Figure 2.4. The 3-Hole Test configuration ......................................................................... 6

Figure 2.5. Example of contact pattern for lightly loaded involute gearing in the 3-Hole 

Test .............................................................................................................................. 8

Figure 3.1. Initial COE comparison with Convoloid technology gearbox ....................... 11

Figure 3.2. Micon 108 involute test gearbox configuration.............................................. 14

Figure 3.3. Early profile inspection trace of Convoloid 13-tooth pinion .......................... 17

Figure 4.1. Conceptual layout of the Micon 108 four-square test .................................... 19

Figure 4.2. The Micon 108 test equipment: The involute gearbox is on the right, the 

Convoloid gearbox is on the left; also visible on the left is the hydraulic cylinder 

(used to introduce loading to the system) ................................................................. 20

Figure 4.3. Convoloid high speed gear mesh no-load contact patch prior to lead 

modification .............................................................................................................. 25

Figure 4.4. Convoloid high speed gear mesh no-load contact patch after lead 

modification; gearing has been assembled into the test housing .............................. 26

Figure 4.5. Audio (sound) spectrum at a distance of 10 ft after approximately 300 hours 

of operation ............................................................................................................... 28

Figure 4.6. Convoloid high speed pinion at conclusion of 200 x 200 testing; note that 

although there is extensive micropitting on both sides of the transition zone, it is not 

rapidly progressing (235.8 hours at 200% of the rated load) .................................... 30

Figure 4.7. Involute HS pinion at the conclusion of 200 x 200 testing; note the extensive 

micropitting in the dedendum and at the edges, some micropitting is visible along 

the tip, this wear is progressing (235.8 hours at 200% load) .................................... 30

Figure 4.8. Vibration spectrum for involute gearbox as measured at the high speed 

bearing; value for the involute high speed gear mesh (533 Hz) is shown in the black 

box (Cursor 2); transmitted vibration from the Convoloid gearbox (433 Hz) is listed 

as Cursor 1; secondary peaks are harmonics ............................................................ 32

Figure 4.9. Vibration spectrum for Convoloid gearbox as measured at the high speed 

bearing; value for the Convoloid high speed gear mesh (433 Hz) is shown in as 

(Cursor 1); transmitted vibration from the involute gearbox (533 Hz) is listed as 

Cursor 2 ..................................................................................................................... 33

Figure 4.10. Audio (sound) spectrum at a distance of 10 feet after approximately 300 

hours of operation ..................................................................................................... 33

Figure 4.11. Failed spherical roller bearing in Convoloid gearbox after 763 total test 

hours; note pitted edges of the rollers and deformation of the cage; metallic particles 

are present at the bottom of the housing bore ........................................................... 34

Figure 4.12. Pieces of the failed bearing ........................................................................... 34
xii

Figure 4.13. Bearing temperature chart for the failed Convoloid intermediate bearing; 

note the rise in temperature after 2:18 a.m.; the jagged “saw-tooth” shape is due to 

cycling of the heat exchanger fan ............................................................................. 35

Figure 4.14. Vibration levels for the Convoloid gearbox for the last day of operation; note 

the increase in vibration at approximately 1:00 a.m. ................................................ 36

Figure 4.15. Involute HS pinion at the end of test; wear due to micropitting is extensive 

along one side (235 hours at 200% load; 528 hours at 240% load) .......................... 36

Figure 4.16. Convoloid HS pinion at the end of test; wear due to micropitting is expanded 

slightly during the test, but no progression occurred on the edges 

(235 hours at 200% load; 528 hours at 240% load) .................................................. 37

Figure 4.17. Involute HS gear at the end of test; no wear is readily visible (235 hours at 

200% load; 528 hours at 240% load) ........................................................................ 37

Figure 4.18. Convoloid HS gear at the end of test; some wear due to micropitting is 

visible on the dedendum side of the transition zone; a large section of the addendum 

still shows marking dye—indicating too much tip relief in the profile (235 hours at 

200% load; 528 hours at 240% load) ........................................................................ 38

Figure 4.19. Involute LS pinion at the end of test; similar to the HS pinion, micropitting 

wear is concentrated along one side and in the dedendum (235 hours at 200% load; 

528 hours at 240% load) ........................................................................................... 38

Figure 4.20. Convoloid LS pinion at the end of test; wear due to micropitting brackets the 

operating surface, except for in the dedendum due to proper tip relief on the mating 

gear (235 hours at 200% load; 528 hours at 240% load) .......................................... 39

Figure 4.21. Involute LS gear at the end of test; no wear is visible (235 hours at 200% 

load; 528 hours at 240% load) .................................................................................. 39

Figure 4.22. Convoloid LS gear at the end of test; slight wear due to micropitting is 

visible in the dedendum (235 hours at 200% load; 528 hours at 240% load) .......... 40

Figure 4.23. Close-up of oil-debris damage on the Convoloid HS pinion; this mark began 

to fade at the next inspection (91.2 hours at 200% load) .......................................... 41

Figure 4.24. Housing stiffness test configuration ............................................................. 43

Figure 4.25. Convoloid HS pinion (77% load, 154 hours) ............................................... 44

Figure 4.26. Convoloid HS pinion (200% load, 235 hours) ............................................. 44

Figure 4.27. Convoloid HS pinion (240% load, 763 hours) ............................................. 45

Figure 4.28. Preliminary layout of 3-Hole Test gearbox .................................................. 48

Figure 6.1. Preliminary layout of 3-Hole Test gearbox .................................................... 50

Figure 6.2. Section view of revised gearbox design ......................................................... 51

Figure 6.3. Final design for the 3-Hole Test gearboxes (section view) ............................ 52

Figure 6.4. Top view of the 3-Hole Test fixture configuration ........................................ 54

Figure 6.5. Micon 108 vibration levels at time of bearing failure .................................... 56

Figure 6.6. One of the chip detectors after a surface fatigue failure ................................. 57

Figure 6.7. Representative involute no-load contact ........................................................ 58

Figure 6.8. Convoloid no-load contact ............................................................................. 58

Figure 6.9. The 3-Hole Test gearbox during final machining .......................................... 59

Figure 6.10. The 3-hole split housing enables easy part changes ..................................... 60

Figure 6.11. The 3-Hole Test fixture ................................................................................ 60

Figure 7.1. Severe scuffing failure due to loss of lubricant .............................................. 63

Figure 7.2. Cutaway of the 3-hole gearbox showing the gear mesh lubricating method . 64
xiii

Figure 7.3. Close-up of Convoloid gear tooth with no micropitting and partial scuffing 

damage due to loss of lubrication ............................................................................. 64

Figure 7.4. Convoloid gearing showing some scuffing and discolored marking dye; the 

left gear was not loaded ............................................................................................ 65

Figure 7.5. No-load contact pattern—involute gear; lube nozzle ..................................... 66

Figure 7.6. Example of the full load (7,000 lb-in) contact—involute gears ..................... 66

Figure 7.7. Involute gearing showing early micropitting; both parts (circled) drive their 

respective mates ........................................................................................................ 67

Figure 7.8. Massive macropitting on pinion tooth flank (Test 3, gearbox 2, pinion 1) .... 67

Figure 7.9. Additional pitting ............................................................................................ 68

Figure 7.10. Further pitting and opposite flank scuffing damage from Test 2 ................. 68

Figure 7.11. Failed involute pinion from Test 4 ............................................................... 69

Figure 7.12. Pinion loaded as a speed reducer .................................................................. 69

Figure 7.13. The other pinion in the same gearbox loaded as a speed increaser and 

showing minimal wear .............................................................................................. 70

Figure 8.1. Representative 750-kW gearbox .................................................................... 72

Figure 8.2. Redesigned low speed (planetary) using Convoloid tooth profiles 

(750 kW gearbox) ..................................................................................................... 74

Figure 9.1. Baseline wind energy with 65-m high tower equals 4,439 MWh per year; 

wind energy generation trade-off 1, tower height increase (COE benefit of 3.5%) . 76

Figure 9.2. Convoloid GB wind energy with 75-m high tower equals 4,598 MWh per 

year; wind energy generation trade-off 1, tower height increase (COE benefit 

of 3.5%)..................................................................................................................... 77

Figure 9.3. Cost of energy comparison for trade-off 1, tower height increase ................. 78

Figure 9.4. Baseline wind energy with 70-m diameter blades is 4,439 MWh per year; 

wind energy trade-off 2, blade diameter increase (COE benefit of 7.5%) ............... 79

Figure 9.5. Convoloid GB Wind Energy with 75-m diameter blades is 4,798 MWh per 

year; wind energy generation trade-off 2. blade diameter increase (COE benefit 

of 7.5%)..................................................................................................................... 79

Figure 9.6. Cost of energy comparison for trade-off 2—blade diameter increase ........... 80

Figure 9.7. Baseline with 70-m diameter blades W.E. with Class 4 wind speed (5.8 m/s) 

equals 4,439 MWh/yr................................................................................................ 81

Figure 9.8. Convoloid GB with 75-m diameter blades W.E. with low wind speed 

(5.539 m/s) equals 4,439 MWh/yr ............................................................................ 81

Figure A.1. Test fixture enclosure .................................................................................... 93

Figure A.2. The 3-Hole Test fixture after paint and initial top milling ............................ 94

Figure A.3. The 3-Hole Test electrical circuit assembly .................................................. 94

Figure A.4. Electrical enclosure........................................................................................ 95

Figure A.5. View inside the test container showing the computer monitoring system and 

polycarbonate view window ..................................................................................... 96

Figure A.6. Initial assembly of 3-Hole Test fixture .......................................................... 96

Figure A.7. The 3-Hole Test fixture showing couplings with alignment covers .............. 97

Figure A.8. Fixed half of new coupling ............................................................................ 97

Figure A.9. Flexible half of new coupling ........................................................................ 98

Figure A.10. Cutaway of the 3-hole gearbox showing the gear mesh 

lubricating method .................................................................................................... 99
xiv

Figure A.11. .................................................................................................................... 100

Figure A.12. Severe scuffing on the Convoloid pinions due to loss of lubricant ........... 100

Figure A.13. The 3-Hole Test gear mesh oil nozzles with proper oil pressure (rotated to 

show flow) .............................................................................................................. 101

Figure A.14. Operational configuration of the 3-Hole Test fixture ................................ 102

Figure A.15. Sample of the monitoring software page (vibration shown) ..................... 102

Figure A.16. 3-Hole Test fixture (top view) ................................................................... 106

Figure A.17. Basic layout of the 3-Hole Test ................................................................. 109

Figure A.18. Representative involute no-load contact .................................................... 110

Figure A.19. Convoloid no-load contact ......................................................................... 110

Figure A.20. No-load contact pattern—involute gear; lube nozzle ................................ 111

Figure A.21. Example of the full load (7,000 lb-in) contact—involute gears ................ 112

Figure A.22. Involute pinion 1—reducer operation (pinion driving gear)—showing 

extensive micropitting on both the addendum and dedendum ................................ 112

Figure A.23. Involute pinion 2—increaser operation (gear drives pinion)—shows a 

smaller amount of micropitting on the pinion (arrow) as well as on the 

mating gear.............................................................................................................. 113

Figure A.24. Involute (1) reducer pinion surface wear (Test 2) ..................................... 114

Figure A.25. Gearbox 1 chip detector after gearing failure; note material crossing 

the gap ..................................................................................................................... 115

Figure A.26. Massive macropitting on pinion tooth ....................................................... 116

Figure A.27. Additional pitting flank (Test 3, gearbox 2, pinion 1) ............................... 116

Figure A.28. Further pitting and opposite flank scuffing damage from Test 2 .............. 117

Figure A.29. Failed involute pinion from Test 4 ............................................................ 118

Figure A.30. Pinion loaded as a speed reducer ............................................................... 118

Figure A.31. The other pinion in the same gearbox loaded as a speed increaser 

showing minimal wear ............................................................................................ 119

Figures A.32. Broken involute gear tooth from Test 5; failure was due to case/core 

separation ................................................................................................................ 120

Figure A.33. Summary of involute gear failures in the 3-Hole Test .............................. 121

Figure A.34. .................................................................................................................... 125

Figure A.35. Contact probability (L9) that matches current failure lives; bending 

L1 life is much higher (for 2X contact cycles) ........................................................ 126

Figure A.36. Calculated L50 contact life and the corresponding bending life (L4.6) ....... 126

Figure O.1. Involute high speed pinion—initial 200% loading; nearly full contact as the 

bluing has been nearly uniformly removed; hard contact is visible on the left side of 

the contact patch; paint marks on the tooth tip are used for identification to verify 

the same tooth was view in subsequent inspections ............................................... 176

Figure O.2. ...................................................................................................................... 177

Figure O.3. Involute high speed pinion—235 hours at 200% load; micropitting 

(dull gray area) in dedendum and at the edge of contact ........................................ 177

Figure O.4. Involute high speed pinion showing extensive micropitting; 763 hours total 

test time (89 x 106

 cycles). (235 hours at 200% load and 1,767 rpm motor speed, 

then the load was increased to 240% and speed increased to 2,050 rpm) .............. 178

Figure O.5 Involute high speed gear at shakedown ........................................................ 178
xv

Figure O.6. Involute high speed gear at shakedown after 763 hours; minimal wear 

is visible .................................................................................................................. 179

Figure O.7. Involute low speed pinion at 200% shakedown; the contact patch reveals a 

consistent contact pattern across the contact flank; note paint marks on teeth tips for 

inspection purposes ................................................................................................. 179

Figure O.8. Involute low speed pinion after 235 hours at 200% rated load; light 

micropitting is visible on all four edges of contact with a heavier concentration in 

the dedendum; faint micropitting also is visible on the grinding peaks (faint lines 

that run horizontal in the active flank) .................................................................... 179

Figure O.9. Involute low speed pinion after 763 total test hours; extensive micropitting is 

visible in the dedendum and at the right side .......................................................... 180

Figure O.10. Close-up of micropitting on the right side of the involute low speed pinion 

after 702 total test hours .......................................................................................... 180

Figure O.11. Involute low speed gear after 200% shakedown ....................................... 181

Figure O.12. Involute low speed gear after 763 total test hours; minimal wear 

is evident ................................................................................................................. 181

Figure O.13. Convoloid high speed pinion after shakedown of 1 hour operation at 200% 

rated load; micropitting on either side of the transition zone appeared during 

approximately 160 hours of testing at 77% rated load; radial marks are due to 

aluminum particles from the heat exchanger that passed through the gear mesh; no 

impression could be felt nor was there any damage to the tooth flank ................... 182

Figure O.14. Convoloid high speed pinion after 235 hours at 200% rated load ............. 182

Figure O.15. Convoloid high speed pinion after 360 total test hours; micropitting around 

the transition zone is broader than after shakedown ............................................... 182

Figure O.16. Convoloid high speed pinion after 538 total test hours; micropitting around 

the transition zone shows minimal progression ...................................................... 183

Figure O.17. Convoloid high speed pinion after 763 total test hours; micropitting zones 

on either side of the transition zone have broadened .............................................. 183

Figures O.18. Convoloid high speed gear at 200% shakedown ..................................... 184

Figure O.19. Convoloid high speed gear at 200% shakedown after 763 total test hours; 

micropitting on the dedendum side of the transition zone corresponds with wear in 

the mating pinion; note the large amount of bluing still visible at the tooth tips ... 184

Figure O.20. Convoloid low speed pinion after shakedown of 1 hour operation at 200% 

rated load; micropitting at the either side of the transition zone appeared during 

approximately 160 hours of testing at 77% rated load ........................................... 184

Figure O.21. Convoloid low speed pinion after 763 total test hours .............................. 185

Figure O.22. Convoloid low speed gear after 200% shakedown .................................... 185

Figure O.23. Convoloid low speed gear after 763 total test hours; light micropitting is 

visible at the edges of the dedendum side of the transition zone; no micropitting is 

evident at the tip, nor is there much bluing ............................................................. 186

Figure O.24. Failed spherical roller bearing in Convoloid gearbox after 763 total test 

hours; note pitted edges of the rollers and deformation of the cage; metallic particles 

are present at the bottom of the housing bore ......................................................... 187

Figure O.25. Pieces of the failed bearing ........................................................................ 187
xvi

Figure O.26. Bearing temperature chart for the failed Convoloid intermediate bearing; 

note the rise in temperature after 2:18 a.m.; the jagged “saw-tooth” shape is due to 

cycling of the heat exchanger fan ........................................................................... 188

Figure O.27. Vibration levels for the Convoloid gearbox for the last day of operation; 

note the increase in vibration at approximately 1:00 a.m. ...................................... 188

Figure P.1. Profile plot of Convoloid high speed pinion showing very small variations; 

the top of the pinion shows the tip relief ................................................................. 189

Figure P.2. Convoloid High speed gear mesh no-load contact patch prior to lead 

modification; gearing has been assembled into the test housing ............................ 190

Figure P.3. Convoloid high speed gear mesh no-load contact patch after lead modification

................................................................................................................................. 190

Figure P.4. HS Involute, 77% shakedown contact .......................................................... 191

Figure P.5. LS Involute, 77% shakedown contact .......................................................... 191

Figure P.6. HS Convoloid, 77% shakedown contact ...................................................... 191

Figure P.7. LS Convoloid, 77% shakedown contact ...................................................... 192

Figure P.8. LS Convoloid gear showing “twist”; note the blue area at the right side of the 

addendum and at the left side of the dedendum (very faint) ................................... 192

Figure P.9. HS Convoloid pinion showing micropitting along the transition zone (63.7 

hours—(approx. 6.8 million cycles) at 77% load) .................................................. 192

Figure P.10. Involute HS pinion with micropitting in the dedendum and edge distress 

(true 200% load after 21.9 hours) ........................................................................... 193

Figure P.11. Involute LS pinion showing edge distress (21.9 hours at 200% load) ....... 193

Figure P.12. Convoloid HS pinion with full face contact, showing transition zone 

micropitting and edge distress (21.9 hours at 200% load) ...................................... 193

Figure P.13. Convoloid HS gear showing slight micropitting (21.9 hours at 200% load)

................................................................................................................................. 194

Figure P.14. Convoloid LS pinion showing micropitting (21.9 hours at 200% load) .... 194

Figure P.15. Convoloid LS gear with faint signs of distress( 21.9 hours at true 

200% load) .............................................................................................................. 194

Figure P.16. Close-up of the Involute HS pinion micropitting (91.2 hours at 

200% load) .............................................................................................................. 195

Figure P.17. Close-up of oil-debris damage on the Convoloid HS pinion; this mark began 

to fade at the next inspection (91.2 hours at 200% load) ........................................ 195

Figure P.18. Micropitting in the Convoloid HS gear; note the extensive area of noncontact at the tip of the teeth, this is due to excessive tip relief applied during 

manufacturing (91.2 hours at 200% load) ............................................................... 195

Figure P.19. Involute HS pinion at the conclusion of the 200% testing; note the extensive 

micropitting in the dedendum and at the edges; some micropitting is visible along 

the tip; this wear is progressing (235.8 hours at 200% load) .................................. 196

Figure P.20. Close-up of involute HS pinion with micropitting (235 hours at 

200% load) .............................................................................................................. 196

Figure P.21. Involute HS gear with no wear at the conclusion of the 200% test 

(235 hours at 200% load) ........................................................................................ 196

Figure P.22. Involute LS pinion at the conclusion of the 200% test showing micropitting 

wear around edges and at grinding peaks (235 hours at 200% load) ...................... 197
xvii

Figure P.23. Involute LS gear at the conclusion of the 200% test showing no wear 

(235 hours at 200% load) ........................................................................................ 197

Figure P.24. Convoloid HS pinion at conclusion of the 200% test. Note that although 

there is extensive micropitting on both sides of the transition zone, it is not rapidly 

progressing (235.8 hours at 200% of the rated load) .............................................. 197

Figure P.25. Convoloid HS gear at the conclusion of the 200% test; micropitting is 

evident in the dedendum side of the transition zone, but the addendum side is free of 

wear (235 hours at 200% load) ............................................................................... 198

Figure P.26. Convoloid LS pinion at the conclusion of the 200% test; micropitting is 

present at all edges except for the bottom of the dedendum (235 hours at 

200% load) .............................................................................................................. 198

Figure P.27. Convoloid LS gear at the conclusion of the 200% test; light micropitting is 

evident around the transition zone and at the root; good tip relief causes the wear 

pattern to fade away at the tip; 235 hours at 200% load ......................................... 198

Figure P.28. Involute HS pinion (235 hours at 200% load; 67 hours at 240% load) ..... 199

Figure P.29. Convoloid HS pinion (235 hours at 200% load; 67 hours at 240% load) .. 199

Figure P.30. Vibration spectrum for involute gearbox as measured at the high speed 

bearing; value for the involute high speed gear mesh (533 Hz) is shown in the black 

box (Cursor 2); transmitted vibration from the Convoloid gearbox (433 Hz) is listed 

as Cursor 1; secondary peaks are harmonics .......................................................... 200

Figure P.31. Vibration spectrum for Convoloid gearbox as measured at the high speed 

bearing; value for the Convoloid high speed gear mesh (433 Hz) is shown in as 

Cursor 1; transmitted vibration from the involute gearbox (533 Hz) is listed as 

Cursor 2 ................................................................................................................... 200

Figure P.32. Audio (sound) spectrum at a distance of 10 ft after approximately 300 hours 

of operation ............................................................................................................. 201

Figure P.33. Involute HS pinion showing progression of micropitting (235 hours at 

200% load; 355 hours at 240% load) ...................................................................... 201

Figure P.34. Convoloid HS pinion showing minimal progression of micropitting (235 

hours at 200% load; 355 hours at 240% load) vertical streaks are from aluminum 

particles (from the heat exchanger) ......................................................................... 202

Figure P.35. Involute LS pinion showing substantial micropitting at the right edge and in 

the dedendum (235 hours at 200% load; 355 hours at 240% load) ........................ 202

Figure P.36. Convoloid LS pinion also shows minimal progression of micropitting 

(235 hours at 200% load; 355 hours at 240% load) ................................................ 202

Figure P.37. Involute HS pinion at the end of test; wear due to micropitting is extensive 

along one side (235 hours at 200% load; 528 hours at 240% load) ........................ 203

Figure P.38. Convoloid HS pinion at the end of test; wear due to micropitting is expanded 

slightly during the test, but no progression occurred on the edges 

(235 hours at 200% load; 528 hours at 240% load) ................................................ 203

Figure P.39. Involute HS gear at the end of test; no wear is apparent (235 hours at 

200% load; 528 hours at 240% load) ...................................................................... 204

Figure P.40. Convoloid HS gear at the end of test; some wear due to micropitting is 

visible on the dedendum side of the transition zone; a large section of the addendum 

still shows marking dye indicating too much tip relief in the profile (235 hours at 

200% load; 528 hours at 240% load) ...................................................................... 204
xviii

Figure P.41. Involute LS pinion at the end of test; similar to the HS pinion, micropitting 

wear is concentrated along one side and in the dedendum (235 hours at 200% load; 

528 hours at 240% load) ......................................................................................... 205

Figure P.42. Convoloid LS pinion at the end of test; wear due to micropitting is brackets 

the operating surface except for in the dedendum due to proper tip relief on the 

mating gear (235 hours at 200% load; 528 hours at 240% load) ............................ 205

Figure P.43. Involute LS gear at the end of test; no wear is visible (235 hours at 200% 

load; 528 hours at 240% load) ................................................................................ 206

Figure P.44. Convoloid LS gear at the end of test; slight wear due to micropitting is 

visible in the dedendum (235 hours at 200% load; 528 hours at 240% load) ........ 206

Figure R.1. Gearbox total vibration at the beginning of the test—200% load ............... 211

Figure R.2. Gearbox total vibration at the end of the test—200% load .......................... 211

Figure R.3. Vibration spectrum of the Convoloid gearbox; cursor is located at the HS 

pinion fundamental frequency (385 Hz) ................................................................. 212

Figure R.4. Vibration spectrum of the involute gearbox; cursor is located at the 

HS pinion fundamental frequency (470 Hz) ........................................................... 212

Figure U.1. Static load test equipment set-up ................................................................. 218

Figure U.2. Close-up of a hydraulic cylinder in position to apply a separating load ..... 219

Figure U.3. Representative strain/force diagram for the involute housing—HS outside 

position .................................................................................................................... 220

Figure U.4. Representative strain/force diagram for the Convoloid housing—HS inside 

position .................................................................................................................... 220

Figure U.5. HS (outside) strain gage reading under combined load—involute .............. 222

Figure U.6. LS (inside) strain gage reading under combined loading—involute ........... 223

Figure U.7. LS (outside) strain gage reading under combined load—Convoloid .......... 223

Figure V.1. Graphical display of the Convoloid tooth form ........................................... 225
xix

List of Tables

Table 3.1. Wind Turbine Gear Drive System (Low-Speed Pass) ..................................... 13

Table 3.2. Wind Turbine Gear Drive System (High-Speed Pass) .................................... 13

Table 4.1. High Speed Involute Gear Set ......................................................................... 21

Table 4.2. Low Speed Involute Gear Set .......................................................................... 22

Table 4.3. High Speed Convoloid Gear Set ...................................................................... 22

Table 4.4. Low Speed Convoloid Gear Set ....................................................................... 22

Table 4.5. Load/Power Summary for Micon 108 Test Original Rating 108 kW (145 hp), 

1,200 rpm at Generator ............................................................................................. 41

Table 6.1. Involute Gear Set ............................................................................................. 52

Table 6.2. Convoloid Gear Set .......................................................................................... 52

Table 6.3. 3-Hole Test Component Naming Specifications ............................................. 55

Table 8.1. Comparison Between an Existing 750-kW Planetary Stage and the 

Convoloid Design ..................................................................................................... 73

Table 9.1. Summary of Wind Energy Generation Options ............................................... 82

Table J.1. Comparison of the Specific Sliding Velocities for Convoloid Versus Involute 

Gear Sets Used in the Micon 108 Wind Turbine Gearbox ..................................... 160

Table J.2. Comparison of the Absolute Sliding Velocities for Convoloid Versus 

Involute Gear Sets Used in the Micon 108 Wind Turbine Gearbox ....................... 161

Table R.1. Convoloid Frequency Spikes and Their Amplitudes .................................... 209

Table R.2. Involute Frequency Spikes and Their Amplitudes (a frequency spike also 

occurred at 88.5 which is a harmonic of the high speed shaft speed) ..................... 210

Table R.3. Motor Frequency Spikes and Their Amplitudes (a frequency spike parallel to 

the shaft occurred at 1410-1414 Hz with an amplitude of 302 x 10-3 G.O-P) ....... 210

Table T.1. Involute Non–Gear Inefficiency Losses ........................................................ 215

Table T.2. Convoloid Non–Gear Inefficiency Losses .................................................... 216
1

1. Introduction

1.1. Background

1.1.1. Low Wind Speed Technology

The U.S. Department of Energy Wind Energy Research Program has begun to develop wind technology 

that will enable wind systems to compete in regions that experience low wind speeds. The sites targeted 

by this effort have annual average wind speeds of 5.8 m/s, measured at 10-m height. Such sites are 

abundant in the United States, and would increase by twenty-fold the available land area which can be 

economically developed. The stated program goal is to reduce the cost of electricity from large wind 

systems in Class 4 winds to 3 cents per kWh for onshore systems or 5 cents per kWh for offshore systems, 

by the year 2012. A three-element approach has been initiated and consists of (1) concept design, (2) 

component development, and (3) system development. This work builds upon previous activities under 

the WindPACT project, the Next Generation Turbine project, and Phase I of the Low Wind Speed 

Turbine (LWST) project. If successful, DOE estimates that this new technology could result in 35 GW to 

45 GW of additional wind capacity being installed by 2020. 

1.1.1.1 Concept Design 

Convoloid gear shapes are a novel new design approach that are expected to provide gears that 

have increased load-carrying capacity but which fit within the same gear envelope of an existing 

gear. (See Appendix V.)This final report includes results from the 100-kW scale testing and 

comparisons to predicted results for a 100-kW design. It describes the design studies conducted, 

including evaluation of the baseline 100-kW gear design loads, predicted lifetime, and other 

appropriate operating parameters. It includes a description of the new gear 100-kW design and 

predicted performance. The report also describes the design studies conducted to scale the 100-kW 

subscale testing and analysis to a MW-class gearbox design, including evaluation of baseline MWclass gear-design loads, predicted lifetime, and other appropriate operating parameters, and a 

comparison to the new full-scale Convoloid design. It explains how the proposed full-scale 

concept is expected to achieve the COE objective for a MW-class system and describes additional 

phases necessary to bring the proposed technology to market. The subcontractor will conduct a 

final review meeting at which the contents of the final report are presented to NREL reviewers.

1.1.1.2 Component Development 

The component development element required the subcontractor to examine and refine the 

proposed technical approach to achieving the LWST cost of energy (COE) objective. The focus of 

this effort is the development of more cost-effective gearing for speed increasers in wind turbines. 

The principal subcontract deliverable is this final report describing the testing and scale-up 

engineering of a Convoloid gear set for a sub-megawatt-scale gearbox for an existing wind 

turbine. The work consists of detailed shop and field testing of a standard 100-kW scale gearbox 

using involute gears and a modified 100-kW scale gearbox using a new Convoloid gear set to 

accurately determine and compare the tooth stresses and operating conditions. Based on the test 

results, the next step is to design a replacement gear set for a larger wind turbine gearbox design. 

The larger gearbox will be 500-kW or greater and be representative of designs used in modern 

megawatt-class machines.
2

1.1.1.3 System Development

In June 2005, Genesis Partners LP was awarded contract number ZAM-5-33200-12 under the U.S. 

Department of Energy Wind Energy Research Program, to examine and develop more costeffective Convoloid gearing and then relate these savings to LWST COE.

1.1.2. Organization

The organization of the project and its personnel is shown in Figure 1.1. 

Figure 1.1. Organization of project personnel

1.2. Objectives

The objective of this study is to develop a Convoloid gear design and to demonstrate its potential to 

reduce the size and weight of gearboxes compared to a baseline involute gear design. This improved gear 

design is quantified in terms of a decrease in capital cost, reduction in gearbox size, and potential increase 

in performance or efficiency.

1.3. Scope

This report describes the approach and rationale used to reach the objectives of the study and the 

organization of the work between Genesis and its subcontractors. It presents the work plan that was 

followed and describes the results of each task. Detailed results of all aspects of the calculations of the 

cost of energy are presented, together with explanations of the appraisal system used to select the most 

promising configurations. A set of optimized configurations is identified and compared with the initial 

baseline design. 

Project Manager

Barney Berlinger

Managing Partner, Genesis 

Partners LP

Genesis Partners LP

V.P., Engineering—Allen 

Williston

Chief Technical Officer—John 

Colbourne

Technical Consultants

Don McVittie (Gear Engineers)

Andy Milburn (Milburn 

Engineering)

Test Consultant

Dick Meredith (DC Energy)
3

2. Methodology—Common to Both Micon 108 and 3-Hole Tests 

2.1. Work Plan

The project was divided into six tasks that reflect the objectives and organization.

• Task 1 involved preparation of comparisons of an enhanced Convoloid gearbox with the existing 

involute-based gearbox for a 108-kW wind turbine. These comparisons addressed stresses and 

operation and maintenance (O&M) costs as well as COE reductions between the two. The task led 

to a kick-off meeting at which Genesis presented its intended approach to NREL. 

• Task 2 focused on specific aspects of the design process leading to testing. Investigations into 

manufacturing methods and sensitivities, gearing inspection, definition of gear failure, and 

component analysis lead directly to the test load values as defined in the test protocol.

• Task 3 involved executing the test protocol for involute and Convoloid gear sets in two different 

test stands.

• Task 4 performed a detailed analysis. 

• Task 5 delineated field testing; the enhancement of the testing in Task 3 precluded field testing for 

this phase of the project. 

• Task 6 created a detailed design of scaled-up gearbox using Convoloid technology and COE 

savings estimates. 

2.1.1. Procedure

The project goal was to develop a Convoloid gear design and to demonstrate its potential to reduce the 

size and weight of gearboxes compared to a baseline involute gear design. This improved gear design 

resulted in decreased capital costs through a reduction in size while potentially increasing performance 

and gearbox life. Development of an appropriate gear design encompassed consideration of traditional 

gear design and rating factors as well as manufacturing methods. Further, a method for checking the gear 

teeth of a manufactured part was needed to insure that the tested geometry was true to theory. Before

testing began, test criteria for load and failure definitions was established and an analysis of all gearbox 

components was performed to assure that no unexpected failures occurred. 

Initial testing involved using the existing configuration for two Micon 108 gearboxes and replacing the 

gearing with optimized involute and Convoloid gearing of the same ratio. Subsequent testing followed 

with a “3-Hole Test” that was not dependent upon the involute and Convoloid gearing having the same 

ratio. An optimized design of the involute gearing was prepared by Don McVittie (Gear Engineers). The 

Convoloid design was developed through an algorithm developed by John Colbourne (Genesis Partners). 

Analysis of all test components at the anticipated test loads was necessary to assure survival during the 

test life span. A summary of Mr. McVittie’s recommendations is included as Appendix Q.

The gearboxes were mounted to a back-to-back test stand (Figure 2.1.) that allowed a great amount of 

torque to be circulated through the gearboxes yet minimized the true power usage to only the amount 

dissipated by system inefficiencies. This method of torque application ensures that both gearboxes are 

subject to the same load. 
4

Figure 2.1. Micon 108 back-to-back test stand arrangement

The realm of operation for wind turbine gearboxes involves extended operation, therefore, a fatigue 

failure mode was desired during testing. Proper selection of the applied torque was necessary to assure a 

contact fatigue failure on the gearing. If the applied load is too great, then tooth breakage could result—

potentially damaging the tested component as well as the test equipment. If the load is too small, then the 

time-to-failure of the test is prohibitive. It was necessary to find a balance between these extremes.

It also was important to constrain as many variables not related to the gearing geometry as possible. These 

constraints included, in part: gearing material and heat treatment, loads, and oil. The lubrication is of key 

importance. A high test load dictated that an assured supply of cool oil was available for the bearings and 

gears. Cross-contamination also must be prevented. Therefore, each gearbox was supported by its own oil 

reservoir, pump, filter, and heat exchanger. Instrumentation for both gearboxes was similar and included 

bearing and lubricant temperatures at various points, lubricant cleanliness, and vibrations in three axes. 

Additional signals tracked circulated torque, motor amperage, and speed. 

Modifications to the Convoloid tooth profile were expected to account for the realities of operating under 

high contact stresses. Although theory might be able to estimate an ideal fatigue life, real-world 

considerations of lubricant film thickness, deflections under load, and manufacturing tolerances were 

expected to require adjustments to the tooth profile. Upon confirmation of these adjustments, full testing 

was possible. Data representing actual surface fatigue failure test results must be compared to theoretical 

estimates to establish a confidence level for the technology.

A determination of the material cost savings, was conducted on a 750-kW wind turbine gearbox. Using 

conservative estimates, these savings were extrapolated to a 1.5-MW size turbine. In addition to cost 

savings, the extended life of the gearing components was determined through reduced stresses in the 

Convoloid gear teeth. The COE analyses of gearboxes using Convoloid technology were developed to 

emphasize the potential impact on the industry. 

2.1.2. Gear Design Methodology

The following rules were observed in the gear-design process. 

• Tooth size and geometry was optimized for the intended test loads. 

• Ratings were based on current American Gear Manufacturers Association (AGMA) standards.
5

• Gearing material was identical for both involute and Convoloid gearing. 

• Tooth modifications (crown and profile) were made to compensate for deflections of gear teeth, as 

well as of all components, creating a smooth mesh under load. The modifications also were 

designed to address lubrication film thickness needs.

2.1.3. Test Fixture Design Methodology

To test the gearing and avoid failures in the test equipment, the following rules were observed. 

• Design (or, in the instance of the Micon 108, modify) gearbox housings for maximum stiffness 

and bore alignment 

• Include additional safety factors in the design of connecting couplings and instrumentation 

components between gearbox shafts. 

• Over-design non-gear components (e.g., bearings, shaft strength, keys) when possible. 

• Provide a convenient means for inspecting gear teeth.

• Verify measurement of test loads through multiple means.

• Assure lubrication and include automated warning and/or test shutdown system. 

• Make visual verification of instrumentation available throughout the test.

Figure 2.2. Micon 108 test fixture, involute gearbox on the left and Convoloid gearbox

on the right
6

Figure 2.3. Micon 108 test fixture, Convoloid gearbox with torque application

hydraulic cylinder

2.2. Testing

2.2.1. Description

The study included two tests. Each test consisted of two gearboxes mounted in a back-to-back

configuration. Figure 2.4. shows the 3-Hole Test configuration as an example of this method of testing. 

Figure 2.4. The 3-Hole Test configuration

A key advantage of this type of test is the ability to circulate a tremendous amount of torque and power 

through the gearboxes using only a small percentage of motor power.. Torque is introduced by a torque 

actuator located between coupled shafts (as shown above) or by rotating one of the gearboxes with respect 

to the other, the method used for the Micon 108 test. Torque control feedback is gathered through a torque 

transducer that is connected to the remaining driveshaft between the test gearboxes.
7

To prevent cross-contamination of the lubricant, both test configurations were designed with separate oil 

systems that consisted of separate reservoirs, supply pumps, cooling radiators, and filters. These were

instrumented for torque, temperature, oil cleanliness, oil flow, and vibration.

2.2.2. Test Gearbox Component Analysis

As with the gearing, all other components were analyzed in an attempt to prevent any non-gearing failures 

within the equipment. Most components experience far more cycles than do the test gears, therefore, 

additional safety factors were included where possible.

2.2.3. Testing Methods

Testing of the Convoloid gearing for this project was divided into two areas, Convoloid gear tooth profile 

development, and a comparison between Convoloid and involute gearing under accelerated loading 

conditions. From the outset of the project, an understanding that the mathematical theory for the 

Convoloid tooth form yields tremendous advantages, production and operation of that profile under realworld conditions would require minute modifications. 

A key interest during these tests was the development of an oil-film thickness sufficient to prevent metalto-metal contact of the gear teeth. During light operation, oil coating the gear teeth completely separates 

the two metal surfaces. As loads increase, surface pressures increase and cause the oil film to be reduced. 

Although the film thickness in the middle of a gear tooth could be very good, a condition exists at the 

edges where oil is “squeezed” out,allowing a metal-to-metal condition due to the reduced film thickness. 

At this interface, premature wear occurs often resulting in a matte-gray surface known as “micropitting.” 

Although this failure mode does not greatly alter the tooth shape, microcracks on the tooth surface could 

precipitate early macropits that shorten the life of a gear tooth. Proper modifications at the edges of 

contact have been shown to minimize or completely alleviate this wear. Determining these modifications 

is the purpose for the initial test protocol.

If the gearing can operate without premature wear, the full fatigue testing can commence. Test protocols 

were developed to ensure that clean, temperature-controlled oil is applied to the bearings and gears. 

Application of loads must be gradual to prevent scuffing between gear teeth. This wear mode is 

characterized by localized welding and then tearing of material at the microscopic level. Degradation of 

the tooth profile occurs rapidly under this failure mode. Prevention is accomplished by assuring proper 

lubricant supply along with the measured increase in tooth loads. 

2.2.4. Test Protocol

Development of test protocols involved preparation of the test fixture and test gearing, preliminary testing 

of the fixture, assurance of oil cleanliness, load application, inspection criteria, determining conditions for 

halting the test, and data collection and storage. Initial operations include marking gear teeth for 

inspection and covering loaded gear flanks with a thin dye for confirmation of tooth contact under load. In 

instances where progressive wear is expected, repeated inspection of the teeth is necessary. Dyeing the 

tooth flanks allows inspection of the tooth contact under varying loads (see Figure 2.5). Proper contact at 

test load is necessary to prevent introduction of excessive stresses in the gear tooth.
8

Figure 2.5. Example of contact pattern for lightly loaded involute gearing in the 3-Hole Test

Prior to beginning a test, all aspects of the test fixture must be verified. In addition to calibration of 

sensors and thermocouples, torque application methods must be confirmed and lubricant flows verified. 

Alarm points within the control equipment also must be confirmed. The final step before applying power 

to the main drive motor is to ensure that contaminants are eliminated from the oil system. After oil 

reaches an acceptable level, gradual application of speed and torque allow any trapped particles to be 

removed. Once the test fixture is ready for full-load testing, a gradual method of increasing the torque 

must be followed to prevent premature lubrication failure. 

2.2.5. Definition of Failure

Prior to beginning testing, a specific definition of failure is necessary. The intent of the testing programs 

is to develop a fatigue related failure comparison, preferably of a surface contact type. Investigation into 

various failure modes and definitions reveals a widespread range of acceptable (or not acceptable) wear 

on gear teeth. The discussion in Appendix D gives a taste for the types of gear failures and the ambiguity 

surrounding when a gear has “failed.” 

AGMA 925-A03: Effect on Lubrication on Gear Surface Distress states:

Laboratory testing commonly uses a 1% limit on tooth surface area damage as a criterion 

to stop a test. However, for field service applications one should always abide by the 

equipment manufacturer’s recommendations or guidelines for acceptable limits of damage 

to any gear or supporting component. 

The purpose of the test was to replicate, as much as possible, a real-world application and to establish an 

emphatic point of gear failure. Therefore, a value of 1.5% of the active tooth face was chosen. Once 

macropitting reaches this point, the loaded flank is assumed to have “failed.”

2.3. Cost of Energy Analysis

The basis of COE used for this study is the 2002 LWST Baseline COE incorporating a 1.5-MW turbine in 

a Class 4 wind environment. Benefits to the COE are calculated by using Convoloid gearing and the 

resultant reduced material usage. Additional COE advantages due to increased service life and the effect 

on repairs during the service life have not been included. See Appendix A and Section 9 for a more 

complete discussion. 
9

2.4. Convoloid Gear Inspection Methods

Due to the unique nature of Convoloid gearing, no methods existed for verifying gear tooth profiles, nor 

were the cutting tools used to manufacture the gearing available. During this project, some methods were 

developed and are discussed here. As more is understood and equipment software is updated, further 

refinement will be necessary. 

2.4.1. Hob Inspection

All Convoloid profiles are specified with x and y coordinates, usually in the transverse plane. In addition, 

the pressure angle is supplied for a given point. From this information, the profile of a basic cutter and, 

ultimately, the true hob form can be created. 

After the hob profile is determined, information can be sent to a coordinate measuring machine (CMM) 

for inspection. A text file containing CMM moves and inspection coordinates is created from the hob 

profile. Deviations from the ideal location are returned by the CMM and can be plotted. Although the 

inspection method would be enhanced with a precision measuring machine (PMM), the data gathered is 

sufficiently accurate to determine whether the physical specimen (hob) is close to the desired profile.

2.4.2. Gear Inspection

For Convoloid gearing, two of the three basic inspections for gearing (index or spacing and lead) can be 

accomplished with existing gear-inspection equipment. The third measurement—profile—only can be 

performed using machines having the latest software. Even with the software, a secondary routine 

currently is necessary to create the deviation charts familiar to those in the gear industry.

Preparation for inspection for Convoloid gears is similar to that for the hobs. However, data supplied is in 

a simple tabular form. After a gear is inspected, the resulting deviations can be plotted. The amount of 

deviation then can be compared to industry standard values for gear accuracy.

3. Project Initiation

3.1. Initial COE Projection

During the project kick-off meeting held August 5, 2005, a projection of the cost of energy for the Micon 

108 with enhanced Convoloid gearing was presented. The projection was based upon wind-load data 

representative of the full-size test location (near Palm Springs, CA). Harry Halloran (Energy Unlimited) 

supplied a power curve based on existing turbines at this location. The emphasis of the study was to 

reveal the change in COE when using the new Convoloid gear geometry versus the industry standard 

involute technology. 

Costs were based upon a production quantity of 250 pieces. The specific area of cost reduction is due to a 

20% reduction in volume and weight of the gearbox, yielding a savings in material costs. This COE 

projection is not based upon an increase in power capacity, although that later could prove valid. 
10

The following aspects of COE for the new gear configurations are evident. 

• Existing Applications—An existing drive could be rebuilt, maintaining the same interface but 

having a higher capacity and/or life. Older units might benefit greatly from the technology due to 

the savings in maintenance and repair costs. Observations from operators in the industry reflect 

that the cost of repair is a large part of the COE.

• New Installation—Gear drives could be a smaller size but have similar gear-stress levels, saving 

weight on the platform. A balance between reduced size and lower gear stresses (yielding longer 

life) could affect new systems dramatically. For large turbines, a reduction in size and weight of 

the gearbox directly affects other systems (e.g., tower, nacelle).

Cost of energy advantages are directly related to the cost for O&M and repairs. Operations and 

maintenance costs include scheduled maintenance; repair costs generally are unscheduled maintenance, 

typically costing between $7,500 to $10,000 per MW per year. Cost of energy trade-offs are addressed in 

detail in Section 9. A summary sheet of the COE is shown in Figure 3.1. 
11

Figure 3.1. Initial COE comparison with Convoloid technology gearbox

3.2. Project Task Outline

The following is a basic summary of the project task outline. The detailed task outline document is 

included in Appendix C. The original outline was modified in May 2006 to shift testing from the Micon 

108 test to the new test fixture referred to as the 3-Hole Test. These items are included under Task 4.

• Task 1—Initial Cost of Energy Projection

• Task 2—Kick-Off Meeting, Conducted 7-27-05 
12

• Task 3—Gearbox Comparison Test and Protocols 

o 3.1. Refine manufacturing processes

o 3.2. Compare stresses at 200% rated load to ensure that bending strength life far surpasses 

that for surface durability

o 3.3. Calculate the life of both boxes under their respective loads in a surface failure 

mode—define “failure”

o 3.4—Analyze manufacturing sensitivities of Convoloid gearing and their effects on the 

rating factors that should be used

o 3.5—Relate testing protocol to expected loads for a Class-4 wind site 

o 3.6—Perform full analysis of housings, bearings, shafts, keys, and other components to 

determine safety factors of these components under protocol loads and speeds (to preclude 

premature failure) 

o 3.7—Describe planned test protocol 

• Task 4—Testing of Involute and Convoloid Gear Sets

• Task 5—Detailed Analysis of Test Results and Predictions

• Task 6—Field Testing

• Task 7—Detailed Design of Scaled-Up Gearbox Using Convoloid Gear Sets 

3.3. Stress Comparisons

3.3.1. Rating Factors

Methods for rating gearing are based upon determining the maximum stresses of the gear tooth (both 

contact and bending) and relating those stresses to the material allowables. Various factors are used to 

modify the applied or the material allowable stresses to compensate for variations in the tooth geometry 

and function. Two important factors for gear rating are gear speed (velocity factor) and tooth mesh 

alignment (load distribution factor). The determination of equations for calculating these factors is 

reviewed continually within the gear industry, as the physics is very complex. Application of these factors 

to the Convoloid tooth form follows the same principles and also requires continual development as more 

is learned.

3.3.2. Projection of Rating Factors and Stress Levels

Table 3.1 and Table 3.2 show the comparison data for the low-speed and high-speed passes of gearing for 

the Micon 108. The general rules governing this comparison were identical to those for the low-speed 

pass, including the determinations of “limiting input torque,” material, heat treatment, and manufacturing 

sequence. In summary, the rules are as follows.

• Ratio: Very nearly the same for both cases; slight ratio changes were made at final manufacture to 

achieve exact ratio equality so that the 4-square rig operated properly

• Center Distances: Same

• Face Widths: Same

• Helix Angle: Slightly larger for Convoloid 
13

• Module: Heavier (larger size gear teeth) for the Convoloid

• Specific Sliding Values: Convoloid is well within tolerance of the values of the involute pair 

Table 3.1. Wind Turbine Gear Drive System

(Low-Speed Pass)

Low-Speed Pass Optimized Involute Convoloid

N1 16 13

N2 80 65

Center Distance 342.9 mm 342.9 mm

Face Width 127 mm 127 mm

Helix Angle 9.0° 12.3894°

Module 7.0 8.6

Gear Ratio 5.0 5.0

Limiting Input Torque 7,681 Nm 10,221 Nm

Ratio of Convoloid limiting torque to involute limiting torque 1.33

Specific Sliding -1.21 to + 0.826 -0.617 to + 0.738

Table 3.2. Wind Turbine Gear Drive System

(High-Speed Pass)

Low-Speed Pass Optimized Involute Convoloid

N1 15 13

N2 94 81

Center Distance 254.0 mm 254.0 mm

Face Width 79.37 mm 79.37 mm

Helix Angle 11.0° 12.19°

Module 4.57 5.28

Gear Ratio 6.267 6.231

Limiting Input Torque 1,712 Nm 2,595 Nm

Ratio of Convoloid limiting torque to involute limiting torque 1.46

Specific Sliding -1.216 to + 0.815 -0.575 to + 0.698

3.3.3. Project Operation and Maintenance Costs as a Function of Stress

When determining the operating and maintenance costs for a wind turbine system, a significant variation 

can occur due to replacement of the main gearbox. This is especially troubling when repairs are 

unexpected. There is a history of premature gear failures involving involute technology, and most O&M 

calculations include planning for gearbox replacement. 

The advantage of Convoloid gearing is that, for a given size, gear stresses are much less than with samesized involute gearing. Lower stresses mean increased gearing life. Replacing involute gearing with 

Convoloid technology—maintaining the existing gearbox and gearing size—and operating under the rated 

loads yields tremendously longer life for the gearbox. Therefore, O&M costs are reduced due to less 

frequent repair and replacement of the gearbox.

3.4. NEG Micon 108 Gearbox Housing Rebuild

Preparation of the housings for testing required careful machining to keep bearing bores aligned and 

maintain proper bearing setting.
14

3.4.1. Housing Rigidity

Initial testing revealed that the single-piece housing lacked sufficient rigidity to properly set the tapered 

roller bearings on the high speed shaft. To prevent the housing faces from spreading apart, four tapered 

pins were installed through the cover plate. This greatly increased the housing stiffness. Figure 3.2 shows 

the involute test gearbox without the cover plate. Additionally, mounting bolts repeatedly pulled out of 

tapped holes in the housing. 

Figure 3.2. Micon 108 involute test gearbox configuration

3.4.2. Bearing Bores

All bearing bores were bored oversized and sleeved to provide dimensional integrity for the bearings. The 

housings had marginal stock around the bores therefore the machining required great care and precision, 

and only very minimal stock could be removed. 

3.5. Involute Gearbox Design

3.5.1. Involute Gearbox Component Manufacture

Manufacture of the involute components follows standard industry practice. The original low speed shaft 

was used for the test as it only required minor clean-up machining. Bearings caps were modified for 

lubrication piping and bearing temperature probes. The high speed shaft was designed with two 

extensions to facilitate coupling to the test drive motor. Steel used for the gearing (identical for both the 

involute and Convoloid) was selected for optimum chemistry and hardenability. Gear blanking, rough 

hobbing, heat treatment, and finish grinding were completed in accordance with standard operating 

procedures. 

Modifications to the keyed connection in the intermediate shaft were necessary because of increased 

stresses that were to be imposed due to the higher test torques. To prevent a buildup of stress 

concentrations at the gear shoulder (point where the gear slides on the shaft and butts up against the 

pinion shoulder), the keyway was extended slightly into the shoulder and the key was lengthened. The 

shoulder radius was also polished.
15

3.5.2. Involute Gearbox Inspection Procedures

Inspection followed industry practice. Gears were inspected using specialized gear-inspection equipment 

and conformed to AGMA quality 12. Further, no-load contact between the pinions and gears was used to 

confirm proper starting contact of mating gearing. The bearing bores diameter and location in the housing 

were verified using a coordinate measuring machine. Upon assembly, all gearing was checked to confirm 

that all rotated smoothly. 

3.6. Optimized Convoloid Design Gearbox

3.6.1. Convoloid Gearbox Component Manufacture

3.6.1.1. General

As discussed, the design and manufacturing of the Convoloid gearbox, accessories, etc. essentially was 

the same as that used for the involute gearbox. 

3.6.1.2. Blanking

The blanking operations for all Convoloid gearing essentially were the same as for the involute gears.

3.6.1.3. Rough Hobbing

The rough hobbing operation essentially was the same as that used for the involute gearing, except that 

small changes in feeds and speeds of the gear hobbing machine could have been advisable to take into 

account the heavier module of the Convoloid designs. Convoloid hobs are made in the same manner as 

highly altered involute hobs, using computer programs to describe the required Convoloid hob rack 

forms. Hobbing Convoloidstakes place in a manner exactly as with involutes.

3.6.1.4. Heat Treating

The Convoloid gears were carburized, quenched, and tempered using exactly the same process as the 

involutes. The only specified difference was the required effective case depth, which allowed for the 

slightly heavier effective module of the Convoloid pairs. The normal precautions regarding case depth 

versus tip width and subsurface shear considerations are consistent with those considerations for involute 

designs. 

3.6.1.5. Final Tooth Grinding

Most—if not all—gear grinding machine manufacturers utilize software protocols based on basic involute 

parameters with additional subroutines to add specified tip relief, twist correction, crown, and other 

anomalies. To date, it has been necessary to use the existing software to grind Convoloid forms. Certain 

procedures have been developed to overcome some of the apparent differences in grinding totally noninvolute forms using involute-based routines. Wheel dressing techniques, inclusion of tip relief, and 

practices to minimize twist have been developed by Genesis Partners LP.

3.6.1.6. AGMA/ISO Gear Accuracy Standards

The accuracies of lead, profile, spacing, and other important quality standards for Convoloid gears mirror 

that of the involute world. Classification of Convoloid gears therefore takes into account the errors found 

in these parameters and then relate those errors to like errors in involute gears.
16

3.6.2. Convoloid Inspection Methods

Due to the unique nature of Convoloid gearing, no methods existed for verifying gear tooth profiles, nor 

did the cutting tools used to manufacture the gearing. During this project some methods were developed 

and are discussed herein. 

3.6.2.1. Gear Inspection

For Convoloid gearing, two of the three basic inspections for gearing (index or spacing and lead) can be 

accomplished with existing gear-checking equipment. The third measurement, profile, only can be 

performed using machines with the latest software. Even with the software, a secondary routine currently 

is necessary to achieve the deviation charts familiar to those in the gear industry. Figure 3.3 shows an 

early profile chart of the 13-tooth Convoloid pinion from the Micon 108 gearbox. The amount of 

deviation can be compared to industry standard values for gear accuracy. 
17

Figure 3.3. Early profile inspection trace of Convoloid 13-tooth pinion

3.6.2.2. Hob Inspection

Traditional hob inspection techniques often employ a visual comparator with scaled representations of the 

intended tooth form. Although this method can be very accurate, no quantitative data can be easily 
18

gathered, making comparisons with AGMA and ISO accuracy standards difficult. Genesis Partners 

developed an alternative method that has proven accurate and reliable. As with gear profile accuracies, 

Convoloid rack forms can be quantified and matched to AGMA/ISO standards. Although the inspection 

method would be enhanced with a precision measuring machine, the data gathered is sufficiently accurate 

to determine whether the physical specimen (hob) is close to the desired profile.

4. Testing—Micon 108

The focus of this test was to refine the Convoloid tooth form from theory to practical application. Once 

qualified, testing continued with the improved Convoloid gear profile in comparison with current involute 

geometry. Stress calculations, especially with respect to contact stresses, reveal dramatic reductions when 

the Convoloid tooth form is used. When designing a Convoloid gear set with contact stresses approaching 

those of involute gears, larger size (greater pitch) gear teeth could be chosen, increasing the bending 

strength of the gear teeth.

4.1. Description of Test

4.1.1. Gearbox and Test Fixture Configuration

The Micon 108 wind turbine gearbox was chosen for the initial comparison test, due to a readily available 

supply. This drive has seen widespread use in the industry, has an extensive history, and is of a 

convenient torque capacity and size for economic load testing (weighing approximately 1,000 kg). A 

detailed discussion of the test appears in Appendix E. 

The two test gearboxes were rebuilt by The Gear Works (Seattle, WA) to like-new condition, which 

included pinning the housings to increase stiffness. The first drive was assembled with high-quality 

(AGMA Class 12) involute gearing. The second drive employed gearing that used the Convoloid 

technology. All material used for the gearing exceeded AGMA Class 2 material specifications, and was 

surface hardened prior to final tooth grind. The involute geometry could be verified by in-shop measuring 

devices, and the Convoloid geometry required outside confirmation via a precision coordinate-measuring 

machine, provided courtesy of Brown & Sharpe (B&S) (North Kingstown, RI). After the gearing was 

assembled with new bearings into the housings, no-load contact patches were obtained for all gearing. 

Approval of these contact patterns was required prior to execution of the test protocol.

The drives were instrumented to enable monitoring of conditions including (but not limited to) all bearing 

and sump temperatures, housing vibrations, and strain between the housing bores. Each drive was 

outfitted with separate oil delivery systems, including pump, primary filter, secondary filter, and heat 

exchanger. The oil was continuously monitored for particulates.

Load application to the drives was accomplished using the “four-square” method, incorporating a 

hydraulic cylinder to torque the common low speed shaft. A strain gauge at the cylinder and a separate 

torque gauge between the high speed pinions provided two methods for verifying the torque applied to the 

system.

4.1.2. Test Fixture

The two gearboxes were configured in a four-square, back-to-back test arrangement. Lubrication systems 

for each drive were isolated with separate filter and cooling systems. Extensive gearbox data was 

collected by a computerized monitoring system. The test fixture was located at The Gear Works (Seattle, 

WA) and was designed to provide a stable platform and withstand heavy loads. 
19

4.1.2.1. Torque Application

The four-square test method is an efficient way to apply a substantial dynamic load without requiring 

extremely large electrical motors and generators. In this type of test, two drives with the same exact ratio 

(27:1, in this case) are mounted to a common low speed drive shaft. The high speed shafts are coupled 

with a torsionally stiff coupling that allows for a large angular misalignment. The advantage of this setup 

is that instead of requiring a motor that is rated for the full power of the tested gearboxes (456 hp), the 

motor only must supply the power lost due to inefficiencies in the test system (approximately 26 hp). One 

gearbox is held fixed and serves to support the motor. A torsional load is applied to the test gearbox, 

thereby twisting the common low speed shaft. The high-speed coupling must accommodate the resulting 

misalignment.

Figure 4.1. Conceptual layout of the Micon 108 four-square test

As shown in Figure 4.1, the involute gearbox (shown on the right side of Figure 4.1) is fixed via a pinned 

connection to the base structure, and the Convoloid gearbox (shown on the left side) is connected to a 

hydraulic cylinder. When pressure is applied to the cylinder, the Convoloid gearbox is rotated around the 

main shaft (supported by pillow block bearings) introducing a twist on that shaft. By tying the high speed 

shafts together and keeping the involute gearbox fixed, the twist induces a torque to the system.

The Convoloid gearbox rotates around the main drive shaft as load is applied through the hydraulic 

cylinder, therefore the high speed shafts do not remain aligned. Some couplings fail to maintain a constant 

rotational velocity as the angle of the shafts increase, so it was important to select constant velocity (CV) 

couplings for the high speed shafts. Further, to monitor the applied torque, a torque transducer was 

mounted between the two couplings. Figure 4.1 shows a good representation of the arrangement. Not 

shown in Figure 4.1 are the oil sump tanks, filters, and heat exchangers. 
20

To minimize as much high speed shaft misalignment as possible (and therefore reduce any potential speed 

variation), at no load the Convoloid gearbox was positioned with the pinion shafts unaligned. The amount 

of initial misalignment was calculated based upon the expected torsional deflection of the solid low speed 

shaft. Minor adjustments on the fixture resulted in the shafts being nearly aligned once the test load was 

applied. 

Figure 4.2. The Micon 108 test equipment: The involute gearbox is on the right, the Convoloid gearbox is on 

the left; also visible on the left is the hydraulic cylinder (used to introduce loading to the system) 

4.1.2.2. Test Fixture Lubrication

In the test, it was necessary to control as many variables as possible. Assuring that all wear components 

were properly supplied with clean oil was of supreme importance. To prevent cross-contamination, two 

identical systems were developed for the gear drives. Test oil viscosity was chosen to be similar to that 

used in the field: ISO320. 

The lubricant system began with the oil sump, which consisted of three chambers that allow any particles 

in the oil to drop out of suspension, and also minimize foaming and trapped air. A heating element in the 

sump was used to bring the oil to operating temperature. Oil was pumped through a high-capacity 10-

micron filter and a heat exchanger. Therefore, clean oil of a consistent temperature was distributed to all 

bearings in the test gearbox. Individual valves for each bearing allowed manual oil-flow corrections to be 

made to maintain bearing temperatures.
21

Although all oil passes through a filter with a 10-micron element prior to entering a test gearbox, a 

separate kidney-loop circuit with an oil pump and 3-micron filter constantly filtered oil in the sump. Oil 

cleanliness of the sump lubricant was monitored constantly.

4.1.2.3. Instrumentation

A number of test parameters were monitored during the test. A detailed listing of those instrumentation 

points is given in Appendix E. 

4.1.2.3.1. Monitored Values

Of high importance was monitoring bearing temperatures and gearbox vibrations, so that any adverse 

deviations could be documented. If a warning threshold was surpassed, then visible and audible alarms 

were set to facilitate the test’s shutdown. Additionally, analysis of the captured data could be used to 

detail methods for earlier notification of impending gear and bearing failures. 

4.1.2.3.2. Controlled Values

Motor speed and applied torque were controlled via manual inputs. Adjustments to the motor set the 

motor speed. A rheostat was used to establish pressure in the hydraulic cylinder to generate torque with 

manual verification of the desired values. During operation, motor speed remained stable; however, slight 

adjustments to the hydraulic pressure were required to maintain the desired torque. 

4.1.2.4. Alarms

Visual and audible alarms were installed to warn of potential gearbox or test fixture failures. Included in 

Appendix E are specific alarm points. As mentioned, all bearing temperatures were monitored with alarms 

set for temperatures exceeding 190°F. Alarms also were set for vibration and lubricant supply. 

4.2. Test Preparations

4.2.1. Test Gearbox Design

Based upon the existing production gearbox, the test gearboxes maintained a basic design of parallel 

shaft, helical gearing supported by roller bearings. Although gearing was optimized, bearings, shaft 

design, and the one-piece housing were retained. 

4.2.1.1. Gearing Design

A four-square test imposes a constraint that the two gearboxes used in the test must have identical gear 

ratios. The ratio used for the test is near that of the original gearboxes: 27.00:1. Under this constraint, the 

gear sets were optimized to produce the highest ratings. Face widths are the same as with the original 

gearbox. The involute gear sets were rated per AGMA standards (ANSI/AGMA 2001), and the Convoloid 

gear sets were rated using the same theory used by AGMA for stresses and life. The two involute gear sets 

(high speed (HS) and low speed (LS) stages) are summarized below.

Table 4.1. High Speed Involute Gear Set

Pinion Gear

Number of Teeth 16 96

Normal Diametral Pitch (Module) 5.776 (4.397)

Normal Pressure Angle 21.9808°

Helix Angle 10.0°

Center Distance 9.8425 in
22

Table 4.2. Low Speed Involute Gear Set

Pinion Gear

Number of Teeth 16 72

Normal Diametral Pitch (Module) 3.175 (8.000)

Normal Pressure Angle 20.0°

Helix Angle 10.0°

Center Distance 13.9764 in

A full analysis using the AGMA Gear Rating Suite of software for the involute gearing under the 240% 

load is included in Appendix F. The two Convoloid gear sets are listed below. 

Table 4.3. High Speed Convoloid Gear Set

Pinion Gear

Number of Teeth 13 78

Normal Diametral Pitch (Module) 4.73128 (5.3685)

Equivalent Normal Pressure Angle 24.96°

Helix Angle 12.292°

Center Distance 9.8425 in

Table 4.4. Low Speed Convoloid Gear Set

Pinion Gear

Number of Teeth 12 54

Normal Diametral Pitch (Module) 2.44079 (10.4065)

Equivalent Normal Pressure Angle 30.60°

Helix Angle 14.679°

Center Distance 13.9764 in

All gearing was analyzed for deflections to determine the proper amount of tooth modifications necessary 

to achieve a smooth, balanced load across the active tooth flank. This analysis required creating an 

analytical model of the gearing, supporting shafts, bearing stiffness, and housing stiffness. The variability 

of the housing stiffness required modifications of the housing to properly support the gearing. 

An analysis of stresses for the test gearing revealed a dramatic reduction in stress levels for the Convoloid 

gearing versus the involute. When applying the 200% load, the contact stresses for the HS and LS 

involute gearing were 174 ksi and 205 ksi, respectively. This compares with the Convoloid contact 

stresses of 167 ksi and 186 ksi (HS and LS respectively). Similar reductions are evident when comparing 

the bending stresses. The complete report is given in Appendix G. 

The minimum anticipated L1 lives for the involute (low speed gear set) was 2,985 hours; the Convoloid 

minimum life (low speed gear set) was 16,750 hours. A complete discussion is presented in Appendix H. 

4.2.1.2. Bearings

• The bearing configuration for the test gearboxes was the same as for those currently in the field. 

One modification was to increase the size (and therefore the rating) of the high speed bearings. 

• Three different types of bearings were used in the Micon 108 gearboxes, tapered roller bearings 

for the high speed shaft; spherical roller bearings for the intermediate shaft; and cylindrical roller 

bearings for the low speed hollow shaft. The lives of the HS and intermediate bearings are much 
23

shorter than those for the LS bearings. At a test load of 240% rated and 2,050 rpm, the expected 

L10 life of the HS bearings was expected to be 280 hours and the anticipated life for the 

intermediate bearings was 418 hours. 

• Making changes to bearings was not possible during this phase of the project. It was expected that 

supplying a ready supply of cool, clean oil to the bearings would increase their life. A summary of 

all internal components in included in Appendix I. 

4.2.1.3. Housing

• During assembly of gear drive components, it was found that the HS tapered roller bearings would 

not retain the necessary preload setting. This was not understood until preliminary testing 

(conducted in June 2005) revealed unexpected shifts in the gear mesh tooth contact patterns as 

soon as power exceeded 100% of rated load (108 kW). The shifts were especially evident at 150% 

of rated load. This problem was evident in both the involute and Convoloid drives.

• Using a “space frame,” a static test of the involute box indicated that twisting of the housing was 

not a problem. A second test to determine whether the housing was deflecting due to gear axial 

forces revealed excessive deflection (0.007 in to 0.009 in). This deflection was reduced 

dramatically when the cover plate was taper-pinned to the housing, essentially becoming part of 

the load-carrying structure. After this modification, tooth contact pattern shifts were not a cause 

for concern. 

4.2.1.4. Component Analysis

The complete report is included in Appendix I. 

• Verification that individual components are capable of withstanding the test loads is good practice 

and included in the detailed task outline (Appendix C) under Task 3. 

• Key analyses revealed a minimum safety factor of 1.03 (for the involute HS shaft extension) at the 

full 240% test load. Other keys were found to have safety factors much greater than 1.0. 

• Experience has shown that rotating shafts supporting gears are loaded with alternating stresses that 

can cause fatigue failures. A critical area of stress is the intermediate shaft where loads from two 

gears apply; specifically, the shoulder where the gear contacts the pinion. During analysis the 

existing design barely met a 1.0 safety factor in this area. Safety was increased by grinding the 

shoulder radius.

4.2.2. Manufacturing Sensitivities of Convoloid Gearing and Their Effects on Rating 

Factors

The mesh action of the Convoloid tooth form is similar to the involute form in that both exhibit a 

conjugate action. This action creates a smooth transfer of power from one component to the other. 

Because of geometry differences between the Convoloid and the involute tooth forms, however, 

calculation methods for key factors used to rate a gear set are modified. Many rating factors are the same 

and react to manufacturing deviations similarly, but differences exist where factors are based upon toothprofile geometry. Those factors with distinct differences are discussed here. The complete report is 

included in Appendix J.
24

4.2.2.1. Specific Sliding

The relationship of the sliding to rolling velocities for Convoloid gear sets are reduced throughout the 

mesh. This reduced “slip” of one gear tooth with respect to the other is highly desirable, reducing the risk 

of scoring and enhancing film-thickness characteristics.

4.2.2.2. Absolute Sliding Velocities

Values for the sliding velocity are computed from the difference in rolling velocities of the pinion and 

gear. The sliding velocities are slightly greater for the Convoloid; however, the relationship of the sliding 

ratio to the rolling ratio (specific sliding) is smaller. This slight increase in absolute sliding velocity is 

thought not to be detrimental, due to the relative curvature of the Convoloid gearing. 

4.2.2.3. Oil-Film Thickness

Film thickness is directly related to the curvature of the elements in contact and their sliding velocities. In 

the traditional involute tooth form, the two mating parts have a convex curvature (similar to two cylinders 

rolling against each other). This contact exhibits a constantly changing radius of curvature for each 

element, creating a varying film thickness over the tooth profile. In gears having small numbers of teeth, 

this condition causes widely varying sliding velocities. In contrast, the Convoloid profile creates a 

situation where one mating tooth portion is convex and the complementary portion is concave (similar to 

one cylinder rolling inside another) and the two profiles have a constant relative curvature. Thus, the oilfilm thickness can be consistent throughout the profile and can exhibit a capacity for greater film 

thickness. 

4.2.2.4. Center Distance Changes

Within normal center distance tolerances common today on modern CNC machinery, Convoloid gearing 

sensitivity to center distance changes will not affect operation or rating expectations. It should be possible 

for Convoloid gearing to function if the center distance is closer than the calculated value. However, 

distinct limits relating to tooth thickness and the location of the transition zone could limit operation of 

the gear set. The application of Convoloid technology in planetary gear sets where the sun gear floats and 

is in tight contact when no load is present has not been tested. 

4.2.2.5. Lead Crown (Tooth Modification)

Although Convoloid gearing can be crowned in a way very similar to involutes; early gear grinding 

revealed the characteristic “twisted tooth” syndrome. This condition is characterized by gears having 

exact leads at the pitch diameter, but with positive lead readings in the dedendum, and negative lead 

readings in the addendum on the same flank of the same tooth, or vice versa. The greater the amount of 

crown specified for a gear, the worse the condition. Subsequent investigation during manufacturing 

helped minimize this condition.

4.2.2.6. Bearing Capacity

Although the radial and separating forces of Convoloid gearing closely approximate those in same-sized 

involute gearing under the same load/speed spectrums, when Convoloid theory is applied to increase 

power density of a gear drive system to carry a certain load/speed spectrum, gear center distances and 

gear sizes are reduced as compared to the involute design (but maintain the same or lower Hertz and 

bending stresses), thus increasing the loads on the supporting bearing system. The bearing industry 

recently developed processes and procedures to significantly increase the power density of products. It is 

thought that these advancements will economically accommodate these anticipated increased loads, 

providing industry with well-balanced gear drive systems from a stress and reliability standpoint.
25

4.2.3. Gear Manufacturing

Manufacture of Convoloid gearing follows the same processes as with involute gearing. Creation of the 

gear blanks and pre-grind gear cutting follow the same methods. Variations between the gear types occur 

when finish grinding the gear teeth. Specifically, because gear grinding software is designed for involute 

gearing, allowances must be made to “convince” the software to properly grind the Convoloid tooth form. 

Additionally, verification of the Convoloid profile is not as straightforward as for involute gears.

Currently, the best confirmation of proper contact between Convoloid gear teeth is to use a marking 

compound to transfer a no-load contact pattern. Once contact was centered on the tooth top-to-bottom 

(Figure 4.3), lead corrections were used to move the contact towards the middle of the tooth (Figure 4.4). 

Investigations conducted during the second phase of testing show promise in developing similar 

inspection techniques as with involute gearing, enabling this type of check to be a confirmation of the 

method, not the sole means of confirmation. 

4.2.3.1. Grinding Lead Modifications into Convoloid Gearing

When gears are subjected to high loads, lead modifications are normally applied to compensate for 

deflections of the gear tooth and supporting shaft. As the amount of modification increases, machine tool 

software must compensate for “tooth twisting” that occurs on a tooth flank. Tooth twisting is a 

phenomenon that typically occurs when grinding both involute and non-involute forms. It manifests itself, 

for example, as a positive lead error in the addendum of the gear, accurate lead at the pitch line, and 

negative lead error in the dedendum creating a cross lead on the same flank of the same tooth. Machine 

tool manufacturers of gear grinding machines have developed software corrections for situations, such 

that lead errors are greatly reduced and twist is very minimal.

Figure 4.3. Convoloid high speed gear mesh no-load contact patch prior to lead modification
26

Figure 4.4. Convoloid high speed gear mesh no-load contact patch after lead modification; gearing has 

been assembled into the test housing

4.2.3.2. Tip Relief

Selection of the amount of tip relief for Convoloid gear teeth was approximated from the involute 

analysis. The Convoloid gear tooth has a greater stiffness than the equivalent involute, therefore the 

Convoloid tooth will not flex as much under load. This leads to a reduced amount of tip relief being 

required. See Appendix P and Figure P.18. 

4.2.4. Test Protocols

This test method is based on an industry-proven method in which 200% of the rated gearbox load (after 

accounting for generator inefficiencies—219,500 lb-in rated torque; 439,000 lb-in test torque) is applied 

for 200 hours. To increase the number of test cycles, the high speed shaft is driven at the motor speed 

(1,767 rpm) instead of the operational field speed of approximately 1,200 rpm. See Appendix K, 

Appendix L, and Appendix M for a more complete discussion of this test method. In an attempt to reduce 

test time, the test load later was increased to 240% of the rated gearbox load and the motor speed 

increased to 2,050 rpm.

4.2.4.1. Shakedown

The shakedown procedure (reference Shakedown Protocol-r2, dated November 17, 2005)(see

Appendix K) was performed until particulates in the oil were well within an acceptable range 

(approximately ISO 14/11/9 maximum for both oil systems). Application of load revealed no problems, 

and the load was increased in increments to 200%. The wear pattern of both drives gradually increased to 

cover the entire face width as the load was increased, as expected. Although the oil was extremely clean 

(oil patch tests confirmed instrument readings), radial marks appeared on the loaded flanks. These marks 

could not be felt and were deemed inconsequential. 

4.2.4.2. Primary Test Protocol

Execution of the test protocol 200% load test (reference Test Protocol, dated September 7, 2005) (see

Appendix L) was as follows. After a successful shakedown, execution of the test protocol commenced. 

Notably, due to a mathematical error in the torque strain gauge chart (converting from micro-strain to 

input torque), the gearboxes were operated at 77% of rated load for 179 hours. Upon discovery of the 

error, the shakedown procedure was followed up to the proper 200% torque value. The test protocol was 
27

implemented once again at the proper load and was concluded after 235 hours. All notations of load in 

this report refer to the actual percent of load applied. 

4.2.4.3. Continued Test Protocol (Continuation of Test—Addendum A of Test Protocol)

Although the original test protocol was complete, testing was continued following the similar procedures 

but at increased loads (240% of rated load) and speeds (up to 2,700 rpm) until a definitive gear failure 

occurred. Loading was chosen after load analysis on all gearing and components was complete. (See

Appendix M.) 

4.2.4.4. Start-Up Procedure

A specific start-up procedure was followed to ensure proper lubrication and to prevent improper overloads 

(Appendix N).

4.3. Testing

See Appendix O for detailed discussions of the testing. 

4.3.1. Early Test Observations—200% x 200-Hour Test Protocol

Note that the original test load was 77%, and not 200% of rated load as was planned. Much of the early 

test observations were recorded during this time. Initially, the Convoloid gearbox seemed noisier than the 

involute gearbox. Once the loading was increased to the true 200% torque, the noise decreased to less 

than that of the involute. The Convoloid gearing also showed early signs of micropitting around the 

transition zone. As this wear progressed on the high speed pinion, contact improved across the face width. 

About halfway through the test, advancement of the micropitting retarded to a point when minimal 

increases were noticed. Noise from the Convoloid gearbox continued to decrease as the test progressed. 

In contrast, the involute high speed pinion began to show signs of micropitting near the root, and this 

wear progressed until the end of the test. Noise and vibration increases accompanied the resulting 

apparent degradation in tooth profile. Although the oil for both gearboxes was exceptionally clean, metal 

particles were evident in the involute samples taken towards the end of the test.

The Convoloid gearbox did not seem to be negatively affected by the application of increased load. As the 

test progressed, the Convoloid gearing seemed to “wear-in,” as opposed to degrading. After the areas of 

high stress were removed by micropitting, minimal additional wear was noted. The involute gearing 

began to wear and micropitting seemed to progress. It also appears that the profile degraded, which 

caused increased noise.

4.3.1.1. Efficiency

Two methods were used to determine the losses within the test system. From these losses, a composite 

efficiency for both gearboxes can be determined. The drives only vary in the gear geometry, therefore a 

valid assumption can be made that there is a negligible difference in efficiency between the two. Thus, the 

efficiency of each gearbox is assumed to be about half the total losses. The total losses were 

approximately 5.4%, giving an average efficiency of 97.3% for each gearbox. A determination for the 

efficiency for each gear mesh can be approximated by removing the power losses due to non-gearing 

components (bearings and contacting oil seals). Without the losses of these components, the efficiency of 

each gear set is 98.9%. (See Appendix T for a detailed discussion.) 
28

4.3.1.2. Sound Comparison

Due to the different tooth counts, there were distinctly different sounds coming from the two gearboxes. 

The involute gearbox had a higher pitch due to having more teeth in the high speed gear set; In 

comparison, the Convoloid gearbox had a decidedly lower pitch. Based purely on subjective impressions, 

at low load levels the sound level of the Convoloid gearbox was markedly louder than the involute. When

loads were increased, however, the Convoloid gearbox became much quieter than the involute. As testing 

progressed, the noise level of the involute gearbox steadily increased but no increase was noticeable with 

the Convoloid drive. 

Data taken with AM audio-spectrum analyzer confirmed what was heard by operators. Figure 4.5 is a 

screen capture showing spikes at the high-speed mesh frequencies of the two gearboxes.

Figure 4.5. Audio (sound) spectrum at a distance of 10 ft

after approximately 300 hours of operation

4.3.2.1.1. Vibration 

Accelerometers were attached to both gearboxes in an attempt to gather vibration data. One finding of 

significant interest is that the Convoloid total vibration value continually decreased as the test progressed, 

and the involute vibration slightly increased. In comparison with the start of the 200% load test, the 

Convoloid vibration reduced by nearly 10%. Conversely, the involute gearbox experienced nearly a 6% 

increase in vibration.

Prior to ending the test, a spectrum analyzer using a magnetic pick-up gathered information on the various 

frequencies of both gearboxes. The nominal high speed frequency was 29.5 Hz (1,768 rpm / 60 sec/m). 

Using this value, the Convoloid high speed gear frequency was 383 Hz (29.5 x 13 teeth) and the involute 

high speed gear frequency was 471 Hz (29.5 x 16 teeth). For the respective gearboxes, distinct peaks 

occur at the fundamental and subsequent harmonic frequencies. The Convoloid fundamental frequency 
29

was much smaller than the involute fundamental. In the first, second, and third harmonics, however, peak 

levels are comparable. Appendix R provides a detailed discussion of this topic. 

4.3.2.1.2. Oil Cleanliness 

Throughout the test, the ISO cleanliness level for 6 micron (6u) particles in the Convoloid gearbox was 7, 

according to the online particle counter. The ISO cleanliness level for 6u particles in the involute gearbox 

was 11. These cleanliness levels were confirmed by daily patch tests for each gearbox. Patches were made 

into slides and have been preserved. In most of the patches debris was difficult to determine. Towards the 

end of the test, however, a small number of metal particles were observed in samples taken from the 

involute gearbox. These were not deemed detrimental to continued operation. 

4.3.2.1.3. Oil Temperature 

The oil sumps and all individual bearing temperatures were monitored and in no instance did any bearing 

exceed 170°F. The hottest bearing in the Convoloid gearbox was the outside low speed bearing with a 

maximum temperature of 167°F which, in the worst case, was 1°F above its sump temperature. The 

hottest bearing in the involute gearbox was the inside intermediate bearing with a maximum of 163°F 

which was also only 1°F above that sump temperature. Both reservoirs were controlled to a maximum of 

161°F and a minimum for 145°F. Variations in the oil temperature primarily are due to the cycling of the 

heat exchanger cooling fans.

Temperatures for the involute gearbox tended to be lower than the Convoloid due to the high air flow 

(approximately 700 fpm) blowing from the motor’s fan. Experimentation with external fans blowing on 

the Convoloid gearbox caused dramatic drops in oil sump and bearing temperatures. Similarly, when the 

air flow from the motor was blocked temperatures dramatically rose in the involute gearbox. Both 

gearboxes generated enough heat to cycle their respective heat exchanger fans, therefore this variation 

was not deemed detrimental to the test.

4.3.2.1.4. Micropitting

Micropitting is a microscopic removal of material that produces a “frosted” or matte-gray surface. The 

topic of the failure mode of micropitting is debated fervently in the industry, and many respected experts 

in the field have differing views. Few that agree as to the underlying causes or as to the effects of 

micropitting on overall gear tooth life. It is agreed that micropitting occurs, however, and that in some 

instances it can lead to macropitting and tooth breakage. 

In this test, it appears that micropitting began in areas of high stress and at the boundaries of contact. High 

stress can be caused by excessive material in the tooth profile, edge and tip loading due to bending of the 

supporting shaft and the tooth itself, and high points within the contacting surface due to grinding marks 

and surface finish. In normal operation, gear teeth are separated by an oil film caused by an increase in 

lubricant viscosity due to very high elastohydrodynamic pressures. At the boundaries of contact, oil 

pressure abruptly drops causing a much thinner film, potentially allowing metal-to-metal contact. 

Solutions to these gear-design challenges required detailed study and careful manufacture. 

4.3.2.1.5. Increase in Test Load to “True” 200% of Rated Load 

The high speed pinion in the Convoloid gearbox experienced micropitting early in the test when subjected 

to only 77% of the rated load. The wear area directly corresponds with areas of positive—or excess—

material. As the test progressed, the other components also exhibited this wear forming a narrow line 

along either side of the transition zone, at the edges, in the root, and at the tip. Near the transition zone, 
30

this line of wear broadened under the 200% load until after approximately 100 hours of operation. Further 

wear abated and did not seem to spread after this point. The initial micropitting along the edges, root, and 

tip did not continue to grow as the test progressed (see Figure 4.6). 

Figure 4.6. Convoloid high speed pinion at conclusion of 200 x 200 testing; note that although there is 

extensive micropitting on both sides of the transition zone, it is not rapidly progressing (235.8 hours at 

200% of the rated load) 

Figure 4.7. Involute HS pinion at the conclusion of 200 x 200 testing; note the extensive micropitting in the 

dedendum and at the edges, some micropitting is visible along the tip, this wear is progressing (235.8 

hours at 200% load)

The involute high speed pinion developed very light micropitting late in the 77% load test. With the 

increase in load to 200%, wear progressed steadily until the test was stopped. A broad line of wear in the 

dedendum along the entire effective face width, at both ends and, to a lesser extent, at the tip is indicative 

of progressive micropitting (see Figure 4.7). The loss of profile due to micropitting caused a noticeable 
31

increase in noise from the involute gearbox. Note also the faint wear at the top of the grinding marks. No 

indications are apparent to predict any impending catastrophic failures with either gearbox. 

4.3.2.1.6. Post-Test Action

To determine the extent of micropitting around the transition zone in the Convoloid high speed pinion, the 

pinion was removed from the gearbox and sent to the North Kingston, Rhode Island, facility of Brown & 

Sharpe for inspection on a Leitz Precision Measuring Machine. Using a 0.50-mm diameter inspection 

probe, the plotted profile of an untouched portion of the active flank showed excess material (0.0004 to 

0.0007) at the extremes of the transition zone, and extending into the addendum and dedendum of the 

pinion. The plot of a worn portion of the tooth showed distinct areas of material removal around the 

transition zone. 

4.3.2.1.7. Preliminary Conclusions

Profile errors found during the post-test inspection of the Convoloid high speed tooth profile and 

improper tip relief (as seen on the high speed gear in Figure 23) caused greater than calculated stress loads 

across the flank of the gear teeth. In contrast, wear patterns on the involute gear meshes confirm that the 

teeth were manufactured to a high-quality, AGMA 12 accuracy. Variations in the Convoloid tooth form 

do not conform to as high an accuracy grade; therefore this test is a comparison of an “A” involute 

gearbox with a “B” Convoloid gearbox. The Convoloid gearbox has functioned as well as the higher 

precision involute gearbox. 

Profile appears to be very important for the Convoloid gearing. Variations in the profile seem to “heal” or 

repair themselves, however, during a wear-in process. What affect this micropitting has and whether the 

wear will progress to macropitting is unknown. Conversely, once micropitting begins on the involute 

gears the wear appears to be progressive, and it can lead to degradation in tooth profile causing increased 

noise and vibration. A detailed discussion of the analysis of vibrations within the test fixture is given in 

Appendix R. Additional observations by the test supervisor are included in Appendix S. 

Although analysis of the individual efficiencies of the gearboxes is not practical in this test due to existing 

instrumentation, the total amount of power lost (as supplied by the motor) is in accordance with the 

average efficiency of two standard gearboxes. See Appendix S for the derivation of gearbox efficiencies. 

4.3.2. Test Observations—Extended Testing

Having completed the testing protocol (200 hours at 200% load and 50% faster speed) during December 

2005, it was determined that the test results were not definitive. Extensive discussions resolved that the 

test should continue until a failure occurs.

To reduce the length of the test, investigations were conducted to determine whether the applied load and 

the motor speed could be increased. Analysis of the limiting factors (gear tooth life, key stress, and shaft 

strength) confirmed that and increase of load to 240% of the operating torque was viable. Speed of the test 

motor was increased to as much as 2,700 rpm, which increased the number of load cycles occurring in a 

shorter period. An addendum to the test protocol was developed (Appendix M). For reference, note that 

240% of the rated load is 515,000 lb-in torque versus the standard 214,500 lb-in. 

Testing resumed on January 23, 2006, and followed the test protocol addendum. Once the oil and bearing 

temperatures stabilized with a 200% load and 1,768 rpm, the load was increased to 220% and then to 

240%. No abnormal increases in temperature were observed and sound levels did not noticeably increase.
32

Motor speed gradually was increased to apply as many load cycles as possible in the shortest amount of 

time. When the speed reached 2,350 rpm, the rubber coupling cover for one of the constant velocity 

couplings ruptured. Upon consultation with the manufacturer, an operational speed of 2,050 rpm was 

chosen. After the coupling was repaired, testing resumed. 

After 67 hours of operation at the 240% load (302 total test hours), micropitting on the involute high 

speed pinion teeth (Figure P.28) appeared to have spread in the pinion root and along one edge (c.f. Figure 

P.20). Wear on the Convoloid teeth (Figure P.29) did not appear much changed from the end of the 

previous test (c.f. Figure P.24). 

Vibration levels in both gearboxes appear to change as the test progressed. Vibration in the involute 

gearbox gradually increased. Conversely, vibration in the Convoloid gearbox slowly decreased. Spectrum 

analysis of these vibrations reveals a substantial difference in the two gearboxes. High speed mesh 

frequencies at 2,000 rpm are 533 Hz for the involute and 433 Hz for the Convoloid. Figure 4.8 and Figure 

4.9 compare the high speed mesh frequency for the involute and Convoloid respectively. When measured 

at the respective gearbox, the fundamental frequency for the high speed mesh is 19.2 dB higher with the 

involute gearbox. 

Additionally, an audio sound spectrum taken from approximately 6 feet away shows a very distinct 

difference in noise level between the two gearboxes (Figure 4.10). The involute gearbox is 12.8 dB higher 

than the Convoloid gearbox, with the primary noise occurring at the high speed gear mesh frequency. 

Figure 4.8. Vibration spectrum for involute gearbox as measured at the high speed bearing; value for the 

involute high speed gear mesh (533 Hz) is shown in the black box (Cursor 2); transmitted vibration from the 

Convoloid gearbox (433 Hz) is listed as Cursor 1; secondary peaks are harmonics
33

Figure 4.9. Vibration spectrum for Convoloid gearbox as measured at the high speed bearing; value for the 

Convoloid high speed gear mesh (433 Hz) is shown in as (Cursor 1); transmitted vibration from the involute 

gearbox (533 Hz) is listed as Cursor 2

Figure 4.10. Audio (sound) spectrum at a distance of 10 feet after approximately 300 hours of operation
34

4.3.3. Cessation of Test

4.3.3.1. Bearing Failure in Micon 108 Test

On March 1, 2006, noise levels and unexplained fluctuations in an intermediate shaft bearing temperature 

in the Convoloid gearbox prompted the test monitor to shut down the test. Significant damage was present 

on the rollers and cage (Figure 4.11). Pieces of the cage and roller material passed through gear meshes 

and other bearings (Figure 4.12). Significant quantities of metal were recovered in the gearbox sump, oil 

reservoir, and primary filter. A day prior to the bearing failure, a magnet inspection of both gearbox 

sumps revealed no metal particles. After the failure, extensive quantities were found inside the gearbox 

housing. 

Figure 4.11. Failed spherical roller bearing in Convoloid gearbox after 763 total test hours; note pitted 

edges of the rollers and deformation of the cage; metallic particles are present at the bottom of the housing 

bore 

Figure 4.12. Pieces of the failed bearing
35

Analysis of temperature and vibration data taken during the final day of operation, show a distinct 

increase in vibration levels about 8 hours before shutdown (Figure 4.13, Figure 4.14). An observation by 

the operator that was written in the test log at the time of the increased vibration indicates that there was 

“a change in the sound” of the Convoloid gearbox. However, testing continued until a more dramatic 

failure occurred. 

Figure 4.13. Bearing temperature chart for the failed Convoloid intermediate bearing; note the rise in 

temperature after 2:18 a.m.; the jagged “saw-tooth” shape is due to cycling of the heat exchanger fan

140

145

150

155

160

165

170

175

180

11:0612:0913:1214:1515:1816:2117:2418:2719:3020:3321:3622:3923:42

0:451:482:513:544:576:007:038:06

Temperature

Time

CO Int Brg Temp

(Actually CI Intermediate Bearing Temperature)
36

Figure 4.14. Vibration levels for the Convoloid gearbox for the last day of operation; note the increase in 

vibration at approximately 1:00 a.m.

4.3.3.2. Condition of Gearing at End of Test

Figure 4.15. Involute HS pinion at the end of test; wear due to micropitting is extensive along one side

(235 hours at 200% load; 528 hours at 240% load) 

0.0

0.5

1.0

1.5

2.0

2.5

3.0

11:0612:0012:5413:48

14:4215:3616:30

17:2418:1819:1220:0621:0021:5422:4823:42

0:361:302:243:184:125:066:006:547:488:42

Time

CONVOLOID VIBRATION LEVELS

C Total Vibration Cx Vibration Cy Vibration Cz Vibration
37

Figure 4.16. Convoloid HS pinion at the end of test; wear due to micropitting is expanded slightly during 

the test, but no progression occurred on the edges (235 hours at 200% load; 528 hours at 240% load) 

Figure 4.17. Involute HS gear at the end of test; no wear is readily visible (235 hours at 200% load; 

528 hours at 240% load) 
38

Figure 4.18. Convoloid HS gear at the end of test; 

some wear due to micropitting is visible on the dedendum side of the transition zone; a large section of the 

addendum still shows marking dye—indicating too much tip relief in the profile

(235 hours at 200% load; 528 hours at 240% load)

Figure 4.19. Involute LS pinion at the end of test; similar to the HS pinion, micropitting wear is concentrated 

along one side and in the dedendum (235 hours at 200% load; 528 hours at 240% load) 
39

Figure 4.20. Convoloid LS pinion at the end of test; wear due to micropitting brackets the operating 

surface, except for in the dedendum due to proper tip relief on the mating gear

(235 hours at 200% load; 528 hours at 240% load) 

Figure 4.21. Involute LS gear at the end of test; no wear is visible

(235 hours at 200% load; 528 hours at 240% load) 
40

Figure 4.22. Convoloid LS gear at the end of test; slight wear due to micropitting is visible in the dedendum 

(235 hours at 200% load; 528 hours at 240% load) 

4.3.4. Variations During Test

A few situations where variability could have been introduced to the test must be noted. 

4.3.4.1. Incorrect Torque Application

Fortunately, the mathematical error in calculating the applied torque from the Lebow transducer caused 

less (rather than more) torque to be applied. Also, because this test ultimately was to compare the 

Convoloid design with an involute, both gearboxes were loaded equally. Thus the comparison remains 

valid. 

4.3.4.2. Bearing Lubrication Method

Lubrication was supplied to each bearing via a manually operated orifice valves. An operator was 

required to constantly monitor bearing temperatures to make certain proper oil quantities were being 

delivered. Although alarm conditions exist for excessive temperatures, smaller variations could have 

caused early bearing wear should the test have been continued. The failure of the intermediate bearing in 

the Convoloid gearbox appears to have been caused by metal fatigue and was not attributed to lubrication. 

4.3.4.3. Lubrication Filtering Before Heat Exchanger

The presence of “smears” on the active surface of the gear teeth (easily visible in Figure 4.23) indicates 

that some small particles went through the mesh. No damage could be felt due to these marks, and old 

marks tended to fade as the test progressed. Oil flowed through the heat exchangers after the moving 

through filters, therefore it is thought that particles flushed from the exchangers could have caused the 

marks. These particles would have been soft aluminum, which would explain the lack of damage to the 

tooth surface. Any particles also would have caused variations in the orifice valves feeding the bearings.
41

Figure 4.23. Close-up of oil-debris damage on the Convoloid HS pinion;

this mark began to fade at the next inspection (91.2 hours at 200% load) 

4.3.4.4. Inadequate Oil Flow Out of the Test Gearboxes

Operators constantly had to balance the amount of oil flow to prevent overfilling the gearboxes due to an 

inadequate draining of the gearboxes. This was understood early in the test; however, modifying the test 

fixture to raise the gearboxes to provide a smoother flow into the oil reservoirs would have substantially 

delayed testing and increased costs.

4.3.5. Load/Power Summary

Table 4.5 shows the loads, speed at the generator, and power summary for the Micon 108 comparison test. 

Table 4.5. Load/Power Summary for Micon 108 Test

Original Rating 108 kW (145 hp), 1,200 rpm at Generator 

% Rated 

Torque

Speed at Generator Circulated Test Power Test Time

(rpm) (kW) (HP) (hr)

77 1,800 94 126 179

200 1,800 242 324 235

240 2,350 386 518 6

240 2,000 432 579 125

240 2,050 443 594 420

It should be noted that the test loads included system inefficiencies. Measurements were taken which 

indicated a total gear inefficiencies and lubricant churning losses to be about 5.2%. Knowing the 

approximate losses from the involute gearbox through conventionally accepted calculations, it was 

inferred that the total losses through each gearbox—involute and Convoloid—were approximately equal 

at 2.6%. Gear mesh efficiencies were approximately 98.9%. See Appendix T. 

When test speeds were increased to 2,350 rpm (high-speed shaft speed) the CV joint lubrication covers 

(made of rubber) failed, thus a decrease in motor speed to 2,050 rpm was necessary. A majority of the test 
42

time occurred at this speed. After 965 hours of total test time, the spherical roller bearings of the 

intermediate bore in the Convoloid gearbox failed. Damage to nearby gearing caused testing to be 

stopped. 

4.4. Test Results

Testing of the existing gearboxes was halted pending analysis of the bearings and feasibility of correcting 

damage due to debris in the gear mesh. Initial plans for replacing the failed bearing and continuing testing 

were dismissed due to the amount of damage sustained by the Convoloid gear teeth. Any further testing 

would require replacement of the damaged gearing, thus negating any previous test results. It was 

emphasized that completion of a laboratory test that shows a definitive comparison of the two gearing 

types is imperative. This is necessary before any field testing can be started.

A field test also is necessary. Due to design modifications in the field, a change in ratio is required for the 

field test gearboxes. To maintain correlation between the laboratory and field tests, gearboxes used in 

both tests must be of the same configuration and ratio, therefore gearing changes are required. 

Additionally, because the failure mode of the test gearboxes was in the bearings, a new bearing 

configuration is required. It was agreed that turbine gearboxes in the field should have filtered and cooled 

lubrication oil and that the oil should be applied directly to all bearings.

4.5. Post-Test Analysis

4.5.1. Housing Stiffness / Bearing Redesign for Micon 108

The bearings are critical to the updated gearbox configuration. Due to failure of the spherical roller 

bearings in the intermediate shaft of the laboratory test, a new bearing design method was required. 

Before any bearing selection could be conducted, however, a test was conducted on the gearbox housing 

to determine its stiffness at the intermediate bearing location.

Mr. Richard L. Meredith (DC Energy) set up one of the test gearbox housings with a hydraulic ram inside 

to measure deflection based upon an applied load (Figure 4.24). A dial indicator was used for 

measurement. After the equipment was installed, the housing cover was assembled to the gearbox as was 

done with the test (i.e., with a bolted connection and all four taper pins installed). Application of hydraulic 

pressure created a force that is trying to separate, or spread the housing. At the nominal rated load (108 

kW) the deflection due to the resultant gear thrust forces (~3,525 lb thrust load) was approximately 

0.0055 in. As the load was increased to the new design load, deflection increased to approximately 0.017 

in. 

These values greatly exceeded expected values and negated any bearing selection that can employ 

opposed roller bearings for the intermediate shaft. Instead, a double-row bearing design was developed. 

For this type of configuration, a double-row tapered roller bearing is employed on one end of a shaft and a 

cylindrical roller bearing is used at the other end. The tapered roller bearing resists all gearing thrust 

forces, allowing the cylindrical roller bearing to carry only radial load. A similar arrangement is intended 

for the high speed bearings. Investigation into limitations of the low speed cylindrical bearings found that 

the existing bearings are sufficient to properly support expected loads. 
43

Figure 4.24. Housing stiffness test configuration

4.5.2. Convoloid Profile Modification (Tip Relief) Study

Determination of the proper Convoloid tooth profile is critical for minimizing localized stresses and 

premature wear. A full discussion is included in Appendix T. In summary, deflections of Convoloid gear 

teeth appear to be 10% to 20% less than an equivalent involute tooth when subjected to the same loads. 

The exact amount of deflection cannot be assumed due to the approximate nature of the analysis. As a 

comparison, however, this analysis has definitely shown that the Convoloid tooth form is much stiffer 

than the involute.

Inspection of the Convoloid high speed pinion allows for a good discussion on tip relief. Figure 4.25 

shows the pinion after 154 hours operation at 77% of rated load. A little bluing is visible at the tooth tips 

(probably due to the chamfer) and on each end (due to crowning), but contact appears smooth at the tip 

with no signs of wear. When the load was increased to 200%, slight wear appears at the tip (Figure 4.26). 

A further increase to 240% reveals distinct tip wear (Figure 4.27). It therefore appears that the applied tip 

relief could be appropriate for the light load, but it is not enough for a heavier loading. 
44

Figure 4.25. Convoloid HS pinion (77% load, 154 hours) 

Figure 4.26. Convoloid HS pinion (200% load, 235 hours) 
45

Figure 4.27. Convoloid HS pinion (240% load, 763 hours) 

The computer analysis indicate that deflection of the gear teeth appears to be not just a function of the 

load—both tooth size and number of teeth make a difference. Additionally, trying to relate the amount of 

correction to an equivalent involute gear might be difficult, although for larger numbers of teeth the 

relationship could be more consistent. Determination of the proper amount of tip relief necessary to 

counter tooth deflection for Convoloid gears requires additional study. At the time of this writing, using 

an approximate relationship for the amount of tip relief could be related to an equivalent involute. 

4.6. Test Conclusion

4.6.1. Test Limitations

Attempting to qualify the Convoloid tooth form in a production wind turbine gearbox is not cost effective 

due to factors including: 

• Flexibility of housing creates variations that cannot be controlled; 

• Assembly/disassembly of components within the full-scale gearbox with a single-piece housing 

design was laborious and very time consuming; 

• Cost of individual parts was very high; and 

• Size of the test fixture is cost prohibitive for numerous tests.

When test experiences required modifications to the Convoloid profile, the entire test fixture had to be 

disassembled. Subsequently, there was a hesitancy to make multiple changes.

4.6.2. Conclusions

Upon analysis of the failed bearing and the resulting damage to the Convoloid components, it became 

clear that additional testing in this test environment would be cost prohibitive. An alternative to 

conducting profile development on the existing test apparatus is to build a much simpler and less 

expensive test gearbox. This configuration (referred to as the 3-Hole Test) features a pinion driving an 

intermediate gear which in turn is driving another pinion (see Figure 4.28). 
46

Two gearboxes were built (one Convoloid and one involute) and arranged in a back-to-back test 

configuration. An advantage of this setup is that the ratios of the two gearboxes need not be exact because 

the second pinion will have the same number of teeth as the first pinion, giving a relative ratio between 

the input and output shafts of 1.00:1. Gear designs for both can be optimized for test loads, thereby 

offering a true comparison. Additionally, more tests can be conducted (two for each gearbox) giving more 

data points and lessening the chances for statistical aberrations.

The project situation meeting held May 18, 2006, featured a summary of the project to date including 

results of the testing conducted in Seattle, Washington; expected field test configurations and gearbox 

quantities; expected cost of energy analysis including the longer durability of a complete gearbox system 

design and additional energy production allowed by higher-capacity gears and bearings; and discussions 

of the benefits of a 3-hole type test. The 3-Hole Test discussion addressed topics including: 

• Long test runs are required to get a surface durability failure with the Micon 108, 

• Assembly and disassembly of the Micon 108 design is cumbersome and labor intensive, 

• A “3-Hole Test” is smaller and much easier to work with,

• Gear change-outs are much easier due to a split-housing design, 

• The idler gear simulates a planet gear with reverse loading, 

• Cost of the gears is much less than those used in the current test,

• Time to surface durability failures is reasonable due to the smaller size,

• Bearings can be sized to prevent bearing failures, and

• Conducting a number of tests increases reliability of test results with Weibull charts or distribution 

plots. 

The overwhelming sentiment of those attending the meeting was that the 3-Hole Test would be of greatest 

benefit to the industry in that more information would be realized for a specific cost. Following as a close 

second was the COE analysis utilizing site-specific wind-load bin arrays applied to the proposed system 

redesign. Careful budgetary projections will allow for the 3-hole testing program, however the 

engineering, manufacturing, and deployment of the field test units have been suspended. 

Preliminary design work commenced on the 3-Hole Test gearbox upon the conclusion of the situation 

meeting. Design power was limited to 180 hp in an attempt to keep the test gearbox small enough that 

testing time would not be too long. The test should be able to run unattended and incorporates a “dry 

sump,” in which the gearing does not dip into the lubricating oil. All bearings and gear meshes are 

directly supplied with filtered and cooled oil. Figure 4.28 is an early representation of a test gearbox. 

4.6.3. Analysis of Test Results—Micon 108

The 4-square test rig was shut down periodically, sometimes more than once a day, to examine the gear 

teeth for failure examinations. Photos were taken to capture detail; these can be found in Appendix O and 

Appendix P. The following observations were made, taking the entire test run in context.

• Stress concentration factors which have a major bearing on tooth bending stresses were confirmed 

to be very adequate for both the involute gearbox (expected due to the long history of theory 

versus practice for this tooth form) and the Convoloid gearbox. Early in 2003, Genesis Partners LP 
47

completed 4-square testing of Convoloid gearing where bending performance was not satisfactory. 

A comprehensive finite-element analysis (FEA) of the Convoloid form was executed with the 

result of much improved protocols and values for this important gear parameter. This information 

was used in the design of the Convoloid gear pairs for the Micon 108 gearboxes. 

• Tip relief calculations and performance worked very well for the involute gear pairs. This 

performance could be expected for involutes for the same reasons indicated above. The Convoloid 

was another matter, however. The much lower bending stresses of the Convoloid designs—and 

thus lower load intensities—affected a recalculated protocol for “LVR” design protocols using that 

data. The test results were “fair” in this regard. This area is one for future study and development 

in the refinement of Convoloid technology. 

4.6.3.1. Micropitting

The involute gearing showed micropitting early in testing that progressed until the test was shut down. 

Some of this could be due to slight lead (helix angle) variations as well as hard contact in the root of the 

high speed pinion. The Convoloid gearing, especially the high speed gear set, showed early signs of 

micropitting around either side the transition zone. As the test proceeded under the full 240% load, the 

thickness of the micropitting wear (0.020 in to 0.030 in) did not grow after the initial 100 hours or so (see 

Appendix P for photos). This test result was studied carefully—including conferences with fluid 

dynamics experts—and new design approaches were developed for future Convoloid designs. One cause 

for the premature wear was confirmed later with gear inspection findings of excessive material around the 

transition zone. This manufacturing deviation caused high localized contact stresses that resulted in the 

premature wear. Once material was removed, no further expansion of the micropitting occurred. 

Investigations into lubricant interaction with contacting gear tooth flanks have also led to slight 

modifications in the Convoloid tooth form to improve tooth contact under load. 

4.6.3.2. Macropitting

Although for themajority of the test, the applied power was at or above 200% of the rated power, there 

was no evidence of macropitting for either the involute gearing or the Convoloid. 

4.6.3.3. Spherical Bearing Failure

Early in the design stages of both the involute and Convoloid gearboxes, a bearing analysis was made to 

assess the load-carrying capacities of all bearings based on the elevated loads anticipated. These loads 

mainly were due to gear forces in the tangential, radial, and axial directions. The spherical roller bearing 

at the intermediate shaft showed a 3,532-hour life at the 200% rated load level. It was thought that this 

calculated life could be significantly lengthened because all bearings in each gearbox were to have 

individual lubrication lines with individually controlled oil flows to keep temperatures in reasonable 

limits. Temperature sensors also were strategically placed to help this effort.

The results, however, were not anticipated. At 965 hours of accumulated test time, a catastrophic and 

sudden failure of the Convoloid spherical roller bearing occurred. Approximately 25 minutes before 

failure occurred, the test operator noticed a slight rise in the oil temperature of this bearing. Oil flow to 

that bearing was increased and nothing suggested an oncoming failure because the temperature rise was 

small and the oil flow increase appeared to rectify the problem. See Appendix O for a complete 

discussion. 
48

Factors contributing to the failure included: 

• A substantial amount of test time was run at 240% load and not at 200% load; 

• The ability to control the oil temperature to that particular bearing was little help in forestalling the 

failure; and 

• The helix angle of the Convoloid intermediate gear set was slightly greater (14.7° versus 10°) than 

that of the involute gear, therefore thrust forces were greater and, as a result, the Convoloid 

bearing failed before the involute bearing had the chance to fail. 

4.6.3.4. Conclusion

Although both gearboxes performed very well under the high accelerated test loads required for an 

accelerated test, much was learned with both designs. To increase the capacity (and design) of bearings in 

both gearboxes would require extensive redesign, housing machining, gear-design changes, and other 

component redesigns and manufacture. A full treatment of the rationale for deciding to proceed with the 

3-Hole Test is provided in Section 5. 

Figure 4.28. Preliminary layout of 3-Hole Test gearbox
49

5. Change of Test Program Venue

5.1. Meeting of NREL and Genesis Partners’ Personnel

The failure of a spherical bearing in the Convoloid intermediate shaft, described in Section 4, indicates 

some of the cost and time implications of fixing that support system and redesigning all bearings in both 

systems. Such a project would be necessary to continue testing at the levels used prior to the bearing 

failure. On May 18, 2006, a meeting was held in Golden, Colorado, to assess the implications of 

continued testing of the Micon 108 gearboxes versus redesigning a smaller 4-square rig (called a “3-Hole 

Test”) and continuing the Convoloid gearing assessment project using this new approach. 

5.2. Analysis of Change of Test Program Venue

The following criteria can be used to justify continuing the Micon 108 testing.

• Substantial data on the performance of both gearboxes has been gathered. 

• The test rig itself and electronic data–collection system are working well. 

• Continued testing, with additional refinements in gear geometries and component upgrades—

especially bearings—will provide significantly more data in “real-world” wind turbine gearboxes. 

• The Micon 108 design demonstrates that there is freedom in gear interfacing components—

especially bearings—where hefty safety factors can be applied to enhance the long-term stability 

and reliability of the test rig.

6. Project Initiation—3-Hole Test

As with the Micon 108 test, the focus of the 3-Hole Test is to refine the Convoloid tooth form from theory 

to practical application and then qualify the theoretical calculations by developing well-controlled test 

results. During testing, involute geometry is used as a control; this also acts to qualify existing gear-rating 

theory, specifically ANSI/AGMA2001 and ISO6336. 

6.1. Description of Test

The basic test concept is identical for the Micon 108 test. A four-square (back-to-back) arrangement was 

developed using a hydraulic torque actuator for load application and direct lubrication to the gears and 

bearing with a dry-sump in the gearboxes. A simple gearbox design facilitated easy part changeover.

6.1.1. 3-Hole Gearbox Design

The concept for the 3-hole gearbox was to build a much simpler,-less expensive test gearbox that would 

enable both the Convoloid and involute gear sets to be optimized for the applied test loads. As shown in 

Figure 6.1, the original concept includes a pinion driving a gear that in turn drives another pinion. The 

pinions have the same number of teeth, therefore the ultimate ratio is 1:1. Thus, exact tooth counts are not 

necessary and gearing can be optimized without limitations. 
50

Figure 6.1. Preliminary layout of 3-Hole Test gearbox 

An additional advantage to this type of testing is that, in each gearbox, the two gear sets act in different 

manners. If the torque is applied through the left-side pinion (see Figure 6.1), for example, then this 

pinion drives the intermediate gear. As the load passes through the gearbox to the right side, the gear 

drives the right-side pinion. In the gearing industry, much discussion has centered on the difference in 

effects of a “reducer” gear set (pinion drives a gear) versus an “increaser” gear set (gear drives a pinion). 

Specifically, the sliding action between the gear teeth determines how well oil can lubricate between the 

contacting flanks. Some gear-design professionals adamantly recommend different gear geometries for 

reduction gearing versus increaser gearing for this purpose; others in the industry are not convinced of the 

necessity. This configuration tests these theories for the involute gears.

6.1.1.1. Conceptual Development

To keep the size of the test manageable, a gearing center distance of 4.00 in was chosen. As the center 

distance increases, the amount of applied load increases (because load capacity is proportional to the cube 

of the center distance). All gearing are expected to have the same materials and heat treatments as 

production wind turbine (WT) gearboxes (high-quality 8620 steel to AGMA 2001 Class 2 specifications).

Initial difficulties arose when ratings of both gear sets were much lower for the common intermediate gear 

due to reversed bending. Future testing will investigate this loading, however the initial testing was 

intended to qualify the surface durability of the tooth forms and to minimize micropitting. In concept, the 

intermediate gear could be modified to cause the pinions to contact in different areas of a much wider face

width. To simplify lead grinding for crown, however, the intermediate gear was changed to two separate 

pieces (Figure 6.2).
51

Figure 6.2. Section view of revised gearbox design

After experience with improper bearing (and, thus, gear) support in the Micon 108 test, a robust housing 

design using an industry-proven design incorporating the readily available tapered roller bearing was 

chosen. A stiff two-piece design with a split at the bearing bore center lines allowed for easy part changes 

and reassembly. Within the given space of the 4.0-in center distance, as much bearing as possible was 

chosen to be assured of no bearing failures. Lubricating oil is supplied to each bearing and to both sides of 

the gear meshes. The bottom of the gear housing is open, allowing oil to drop into a reservoir. Each 

gearbox has separate lubrication systems.

Analysis of related components and an unexpected failure in the intermediate shaft required changes in 

the intermediate bearing selection and a strengthening of the shaft. The final design is shown in 

Figure 6.3. 
52

Figure 6.3. Final design for the 3-Hole Test gearboxes (section view)

6.1.1.2. Gearing Design

Gearing was designed to achieve expected L1 lives of approximately 200 hours with the original test load 

(6,300 lb-in). When calculating the average lives at this load (L50 > 24,000 hours), it was realized that too 

much time would possibly pass between subsequent tests. An increase in load to 7,000 lb-in reduced the 

average life to approximately 6450 hours (L1 = 11 hours). Additional increases were planned, but early 

gear failures (surface contact, macropitting) suggested maintaining the current loading. The involute gear 

set is summarized below.

Table 6.1. Involute Gear Set

Pinion Gear

Number of Teeth 16 35

Normal Diametral Pitch (Module) 6.000 (3.629)

Normal Pressure Angle 20.0°

Helix Angle 18.6974°

Center Distance 4.00 in

A full analysis using the AGMA Gear Rating Suite of software for the involute gearing under the 

7,000 lb-in test load is included in Appendix A. The calculated lives for the Convoloid gears were much 

longer than with the involute, with an average life of more than 185,000 hours and an L1 life of 320 hours. 

The Convoloid gear set is as follows.

Table 6.2. Convoloid Gear Set

Pinion Gear

Number of Teeth 11 24

Normal Diametral Pitch (Module) 4.8917 (6.1925)

Equivalent Normal Pressure Angle 23.72°

Helix Angle 26.5732°

Center Distance 4.00 in
53

All gearing was analyzed for deflections to determine the tooth modifications necessary to achieve a 

smooth, balanced load across the active tooth flank. This analysis required creating an analytical model of 

the gearing, supporting shafts, bearing stiffness, and housing stiffness. Some modifications to the HS 

pinions and the intermediate shaft were required due to excessive shaft deflections under load.

6.1.1.3. Housing Design

The housing was designed with two halves to make it easier to change out gear components. The base of 

the lower housing half has a large opening so the oil drops freely into the reservoir. The two halves are 

pinned together during machining operations, maintaining bore integrity and alignment during 

changeovers. 

6.1.1.4. Component Analysis

• Due to the high loads being circulated through the gearboxes, analysis of the individual 

components is necessary, specifically keys, bearings, and shaft stresses.

• The original conceptual design featured a single gear that passed the load from one pinion to the 

other. A key was not necessary because no load was passed to the shaft. Upon changing to two 

gears, the key had to carry the entire load. Analysis of the key prompted a redesign of the 

intermediate shaft that increased the gear mounting diameter. The final design yielded an 

appropriate key safety factor for all test loads.

• Bearings were replaced whenever a component required replacement. Most components were 

loaded from both directions (first one flank and then the other), however, so the bearings had to be 

able to withstand twice the number of expected gearing cycles.

• The original 3-hole design incorporated bearings that were as large as possible to assure long life. 

For the larger intermediate bearing, however, the shaft size was diminished and weakened. Shaft 

stress calculations at the bearing shoulder revealed enough capacity for early test loads 

(6,300 lb-in), but when using the higher loads necessary to reduce test time (7,000 lb-in), one of 

the shafts broke. The calculated safety factor at the higher load was less than 1.0.

• A change in bearing (slightly smaller and with lower load capacity) enabled the shaft to be 

strengthened, increasing the safety factor to 1.59. Although the intermediate bearing had a lower 

rating than the original, during testing no wear was found on any of the bearings and no additional 

shaft failures occurred.

6.1.2. Test Equipment and Fixture Design

6.1.2.1. Test Fixture

The purpose of the test fixture is to properly support the gearboxes, supply load to the gearing, rotate the 

gearing, and supply clean lubricant of a specified temperature. During operation, the data gathered to 

monitor the test should include torque application, temperatures, vibrations, and lubricant cleanliness.

During testing, a large amount of torque was applied to the system. This torque acts to “twist” one test 

gearbox with respect to the other. It therefore was imperative that the mounting surface of the fixture be 

stiff enough to maintain gearbox shaft alignment under test loads. Additionally, to create a compact 

testing area the fixture was designed with internal reservoirs to accommodate all the oil required for the 

test. Further, separate reservoirs are included to prevent cross-contamination between the two gearboxes. 

Each reservoir had two baffles to allow particles to settle and to reduce foaming.
54

The intent of the design was to create a platform for mounting not just the gearboxes but the drive motor 

and lubricating pumps. This platform, along with the rest of the test equipment (heat exchangers, filters, 

and computer control system), then fit into a shipping container allowing for unobtrusive placement 

outside of normal manufacturing space. This was necessitated by requests to free up the main test 

platform and reduce noise. 

Lubrication was straightforward, with oil being directly applied to each bearing through the bearing caps 

and to each gear mesh via oil jets. A “dry-sump” design enables all lubricating oil to fall out of the 

gearbox and into the fixture tank. As with the previous test, each gearbox includes a separate lubricant 

system with separate filters and heat exchangers.

The fabricated fixture shown in Figure 6.3 met all design intents. Each lubrication system (one for each 

gearbox) uses two oil pumps. One pump directs oil from the reservoir through a 3-micron filter into a 

distribution manifold that, in turn, feeds oil to each bearing and to both sides of the gear meshes. Oil flow 

is controlled by small orifices in the manifold, thereby ensuring oil flow to all components. A second 

pump circulates oil from the reservoir through a heat exchanger, past a heater, and back into the reservoir, 

thus maintaining a consistent oil temperature.

After initial alignments, the test gearboxes were pinned to the top of the fixture. It was found that part 

changeovers were accomplished more easily if a gearbox was removed from the test fixture. Tapered pins 

provided assurance that the shafts would be aligned when the gearbox was remounted to the fixture. 

Figure 6.4. Top view of the 3-Hole Test fixture configuration

6.1.2.2. Load Application

Similar to the Micon 108 test, rotation was provided by a directly mounted motor. To mount the 

gearboxes to the solid top of the test fixture, a different method was needed to apply torque. By using a 

hydraulic rotary torque actuator with a rotating union, torque could be applied in either direction without 

affecting shaft alignment—allowing use of standard couplings. Being able to change the torque direction 
55

was necessary for testing both flanks of the gear teeth. Torque was monitored using the same Lebow 

transducer from the Micon 108 test. 

A standard 3-phase motor with a variable frequency drive (VFD) was chosen to drive the system. Using 

the design power of 180 hp (at 1,800 rpm) and a total 6% loss in the system, 10.8 hp is required. To apply 

as many load cycles as possible, the test operated at 2,200 rpm—thus requiring about 13.5 hp from the 

motor. Because of the winding losses due to the increase in speed, and to keep motor temperatures down, 

a 30 hp motor was used. 

Operation of this 3-Hole Test allows for two different mesh conditions (a reduction gear set and an 

increaser gear set) in each gearbox (involute and Convoloid). Therefore, four gear meshes were tested at a 

time. Subsequent testing also can be conducted by reversing the direction that torque is applied. Thus, 

sufficiently descriptive and unique annotations were required for each component as well as each tooth 

flank. Referencing Figure 6.4, gearing component designations could follow the descriptions in Table 6.3. 

Table 6.3. 3-Hole Test Component Naming Specifications

Involute Gearbox Convoloid Gearbox

1 Involute pinion 1 5 Convoloid pinion 1

2 Involute gear 1 6 Convoloid gear 1

3 Involute pinion 2 7 Convoloid pinion 2

4 Involute gear 2 8 Convoloid gear 2

As an example, for a given direction of torque, gear sets 5, 6, 3, 4 could act as increasers, and gear sets 1, 

2, 7, and 8 could act as reducers. This changes as the torque reverses directions to gear sets 5, 6, 3, and 4 

acting as reducers, and gear sets 1, 2, 7, and 8 acting as increasers. Determining whether a gear set is 

acting as an increaser or a reducer is based upon the motor direction and which gear tooth flanks are 

loaded. 

6.1.2.3. Instrumentation

Test instrumentation is similar to that of the previous test; however, additional controls for torque 

application were added to facilitate unattended operation. A substantial cost for the previous test was a 

requirement of having an operator monitor the test at all times. This was necessitated by constant 

variations in lubricant flow and fluctuations in the torque control required constant attention. 

For the 3-Hole Test, alarm points were inserted in the control software to enable the system to carefully 

shut down when appropriate. Key among these shutdown parameters were high bearing temperature, high 

vibrations, and detected particles in oil flowing out of the gearboxes. Further, external conditions were 

monitored (test-fixture compartment ambient temperature, computer compartment ambient temperature, 

electrical power, smoke detector) so that the system could be automatically shut down in the event of an 

emergency. A detailed list of monitored parameters is given in Appendix A. 

6.1.2.3.1. Monitored Values 

Monitoring bearing temperatures and gearbox vibrations so that any adverse deviations could be captured

was very important. If a warning threshold was surpassed, then visible and audible alarms were triggered 

in addition to the automatic shutdown of the test. As determined from the previous test, temperature could 

not necessarily be considered a dependable variable. Also, total vibration did not reveal a problem until 

major damage occurred. One aspect of vibration did directly correlate to the beginning of damage, 

however (Figure 6.5, see vertical line that corresponds to test documentation made by the operator of a 
56

CONVOLOID VIBRATION LEVELS

0.0

0.5

1.0

1.5

2.0

2.5

3.0

11:0611:5412:4213:3014:1815:0615:5416:4217:3018:1819:0619:5420:4221:3022:1823:0623:54

0:421:302:183:063:544:425:306:187:067:548:42

Time

C Total Vibration Cx Vibration Cy Vibration Cz Vibration

“change in sound” in the test). In one direction the levels were generally small; however, once damage 

started, this very small component of the total vibration dramatically changed. Analysis of the starting 

value of the small vibration component was used to establish an alarm value.

Figure 6.5. Micon 108 vibration levels at time of bearing failure

6.1.2.3.2. Controlled Values 

In contrast with the earlier test, motor speed and applied torque was computer controlled and monitored. 

Any variations in the rheostat setting were adjusted, keeping the system much more consistent than when 

corrected manually. 

6.1.2.4. Alarms

In addition to external visual and audible alarms installed to warn of potential gearbox or test fixture 

failures, e-mail and telephone notifications were added due to the unattended operation of the test fixture. 

Included in Appendix A are specific alarm points. As noted, all bearing temperatures were monitored with 

alarms set for temperatures exceeding 190°F. Alarms also were set for vibration and lubricant supply. 

Another important alarm condition was related to the presence of any metal particles in the oil falling out 

of the test gearbox and back into the reservoir. This oil was funneled past a chip detector (Figure 6.6) 

designed to collect metal particles with a magnet. When particles are present a gap is bridged, triggering 

the alarm.
57

Figure 6.6. One of the chip detectors after a surface fatigue failure

6.2. Test Preparations

6.2.1. Test Equipment

6.2.1.1. Gear Manufacturing

Manufacturing of the gearbox housings, the involute gearing, and the Convoloid gear blanks was 

accomplished with no problems. Twist correction protocols are inherent in most gear tooth grinding 

machine software, and depend on basic involute data to effect the necessary machine/wheel movements to 

mitigate twist. Convoloid forms, being non-involute, do not respond as well to these twist correction 

inputs as do involute tooth forms. Verification of the tooth contact was used to apply slight machine 

profile corrections resulting in good no-load contact. Upon installation of the gear in the test gearbox, and 

painting it with machining dye, the light-load contact patterns were confirmed (Figure 6.7, Figure 6.8). 
58

Figure 6.7. Representative involute no-load contact

Figure 6.8. Convoloid no-load contact

In the ideal case, a no-load contact shows marking centered on the tooth flank, and the edges of the 

marking gradually “feather away” leaving no hard edges. For the Convoloid gearing, there should be an 

empty area in the center of the flank around the transition zone with contact above and below—again with 

feathered edges. The original Micon-108 test had very hard contact around the transition zone. It was in 

this area that micropitting appeared very early in the test and at a relatively low load.

The intent of the initial testing is to determine the proper tooth geometry to minimize or eliminate 

micropitting around the transition zone. The test is designed to test one flank at a time for each gear 

element, therefore if good contact can be achieved on one flank then the hardened test parts can be ground 

and the test begun with no detriment to viability of the test. Whether micropitting appears with the 
59

updated geometry should be evident early in testing. Should micropitting occur, the gearing geometry is

evaluated and the parts are reground. When a test can be run without premature wear (micropitting), then 

long-term testing can begin. 

6.2.1.2. 3-Hole Gearbox Housing

Manufacture of the gearbox housings was straightforward. The design consisted of fabricated steel upper 

and lower housing halves to provide exceptional bearing support. The housings (Figure 6.9) were stress 

relieved to prevent any unexpected distortions during testing. The two halves were pinned and bolted 

prior to final machining to assure bearing alignment. Housing stiffness for this test is much higher than in 

the previous test. In contrast to the one-piece housings for the Micon-108 test, these split housings allow 

easy replacement of gearbox components (Figure 6.10).

Figure 6.9. The 3-Hole Test gearbox during final machining

6.2.1.3. Test Fixture

The test fixture (Figure 6.11) consists of the two test gearboxes, the main test fixture (with drive motor, 

Lebow torque transducer, rotary torque actuator, and lubricating pumps), electrical control assembly, and 

the computer control. 
60

Figure 6.10. The 3-hole split housing enables easy part changes

Figure 6.11. The 3-Hole Test fixture 
61

The test fixture base was designed to hold all the oil for the test in separate baffled reservoirs. It was 

necessary to have a separate lubrication system for each gearbox to prevent cross-contamination between 

the gearboxes. Two separate pumps were fitted for each system—one for circulating oil taken from the 

reservoir through the heat exchanger and heater; and, one for supplying filtered oil to the oil manifold for 

distribution to each bearing and to both sides of both gear meshes. A detailed discussion of the test fixture 

configuration and shakedown is included in Appendix A. 

6.2.1.4. Testing Shutdown Criteria

The occurrence of a fatigue failure is an obvious reason to halt testing. During testing, however, other 

conditions can occur that require ceasing operation immediately. Any change in sound, vibration, oil 

cleanliness, or any other abnormal operation is acceptable cause to shut down the operation of the load 

cell and motor (lubricant pumps should remain engaged until all motion is stopped). A thorough 

inspection of all components should be conducted before continuing the test, and all variations should be 

noted in the test notebook. A list of computer monitored conditions follows; however, there can be 

additions or deletions as needed. Each of these conditions has warning levels and shutdown levels 

enabling the test to run unattended. 

• Loss of oil pressure (pumps failed)

• Oil differential pressure too high across filters (filters clogged)

• Chip detector goes low (ferrous material in the oil)

• Bearing temps are greater than 190 o F 

• Oil flow too low

• Oil pressure too high at manifold (manifold orifices clogged) 

• Vibration over limits in any of three planes

• Particle counters indicate high particulate count 

• Torque loss 

• Power loss to test container

• Bulk oil over temperature

• RPM of drive system out of test limits

• Visual observations from the webcam indicate shutdown is needed 

6.2.2. Test Protocols

6.2.6.1. Micropitting Test Protocol

Reference: Micropitting Test Protocol, dated April 9, 2007 (see Appendix A). 

Initial testing in the 3-hole gearboxes was intended to qualify profile corrections in the Convoloid tooth 

form. This testing is in preparation for the main extended comparison test. Previous test gears developed 

micropitting early in the test cycle. Although this failure mode is not in itself catastrophic, it can lead to 

macropitting and, in extended operation, to complete failure of the gear. Before extended testing can 

occur the micropitting phenomena must be solved. 
62

6.2.6.2. Extended Comparison Test Protocol

If premature micropitting wear is prevented, then extended testing for comparisons of the surface 

durability is possible. The expected failure mode is surface fatigue on the contacting gear tooth flanks 

which results in macropitting. A failure for this test was defined as follows.

The definition of failure for this test is the development of an area of macropitting over 

more than 1.5% of the contact face width i.e. active tooth flank. The two gear sets (involute 

and Convoloid) are of a different module (tooth size), therefore the actual area varies 

slightly. The amount of area that is assumed to be a macropitting (fatigue) failure for the 

involute is an area of 0.0070 in2 (4.52 mm2

) which can be described by an area bounded by 

a circle of an approximate 0.10 in (2.4 mm) diameter. For the Convoloid gearing, the area 

is slightly larger {0.0117 in2 (6.55 mm2)} with a diameter of 0.12 in (3.1 mm).

Extended Comparison Test Protocol, dated April 9, 2007 (see Appendix A). 

Torque values for testing originally were expected to be 7,880 lb-in, however this was reduced to 

6,300 lb-in. After initial operation at this value, loading was increased to 7,000 lb-in where it remained for 

the remainder of the test.

7. Testing—3-Hole Test

7.1. Testing

Appendix A provides detailed discussions of the testing. To date, the majority of testing occurred during 

shakedown of the test fixture. As such, some gear failures could not be attributed to traditional surface 

fatigue (the expected failure mode). Also, most testing has involved two involute gear sets (versus one 

involute and one Convoloid) due to manufacturing delays and the micropitting protocol. The number of 

cycles and load amount are being logged for each loaded flank, so all test cycles can be tracked. All 

gearing failures have been documented to help generate a baseline for future Convoloid comparisons. The 

definition of gear surface fatigue failure is discussed in Section 6.2.6.2 (above). 

7.1.1. Early Test Observations—Micropitting Test

Before starting any extended testing, the premature micropitting wear that appeared on the Convoloid 

gear teeth in the Micon 108 test had to be eliminated. In the earlier test, micropitting occurred early at 

lower test loads, therefore initial testing was focused on proving the tooth profile corrections, and 

resolved this situation. The Micropitting Test Protocol is included in Appendix A. 

Execution of the protocol began during shakedown of the test fixture under constant supervision. When 

the test fixture was up to temperature, low oil pressure in the distribution manifold that feeds oil to the 

bearings and gear meshes prompted inspection of the lube pump bypass valve leading to a loss of 

lubricant to the Convoloid gearbox. Although operation was stopped, severe scuffing damaged the gear 

teeth (Figure 7.1).
63

Figure 7.1. Severe scuffing failure due to loss of lubricant

Some teeth suffered worse damage than others due to how the load was applied, the direction of gearing 

rotation, and how the oil was supplied to the tooth mesh. If the gear mesh was loaded (as in Figure 7.2), 

then loss of oil pressure would starve the contact area of lubrication. No oil would could flow upward,

against gravity. The upper oil nozzle only acts to cool the gear teeth after emerging from the mesh. 

However, the other gear mesh received a little more oil, due to gravity allowing oil to drip into the mesh. 

When oil failed to lubricate the gear teeth, localized temperatures raised dramatically causing minute 

welding and tearing of the tooth flanks. Fortunately, some information can be gleaned from the 

undamaged portions of some teeth in the gravity lubricated mesh. 
64

Figure 7.2. Cutaway of the 3-hole gearbox showing the gear mesh lubricating method

Figure 7.3. Close-up of Convoloid gear tooth with no micropitting and partial scuffing 

damage due to loss of lubrication

The loaded tooth flanks show some scuffing damage (Figure 7.3, Figure 7.4), but other areas still show 

grinding marks. Due to early micropitting around the transition zone in the Micon 108, slight 

modifications were made to the tooth profile algorithm in this area. Inspections of the undamaged portion 

of the gear teeth do not show any early wear from micropitting around the transition zone giving a 

(preliminary) verification to the changes. Enough torque (6,300 lb-in) had been applied to generate wear 

if similar geometry conditions to the Micon 108 test had existed. However, not enough test time passed to 

allow for a definitive assessment. Only one full Convoloid gear set had been finished ground, pending 

qualification of the final tooth profile geometry. Continued micropitting testing on the Convoloid gearing 

therefore had to be postponed until new parts were ground. 
65

Figure 7.4, Convoloid gearing showing some scuffing and discolored marking dye; the left gear was not 

loaded 

7.1.2. Test Observations—Extended Testing

Operation of the test fixture was required to properly shake down the equipment. Because the calculated 

average life was greater than 8,000 hours, applying load cycles whenever possible was thought to help 

achieve usable results as soon as practical. Therefore involute gearing was installed in gearbox 2. All test 

cycles were documented and it was intended that any information gathered from the involute gearing 

would act as a baseline for the Convoloid testing. Further comparisons of the involute gearing with 

gearing-life calculations based on industry standards will be meaningful to the industry. 

As with previous testing, the no-load and full-load contact between all gearing were documented to verify 

proper gear mesh (Figure 7.5, Figure 7.6). The slight presence of marking dye at the corners and in the 

root and tip indicate that there is no hard contact in the tooth mesh that would lead to premature wear. 
66

Figure 7.5. No-load contact pattern—involute gear; lube nozzle

Figure 7.6. Example of the full load (7,000 lb-in) contact—involute gears

To reduce the average expected test life to less than a year of constant operation (24 hours per day, 7 days 

per week) test loading was increased from the original 6,300 lb-in to 7,000 lb-in. Additional increases in 

torque were not applied due to development of micropitting on some of the gear teeth (Figure 7.7). There 

was a distinct difference in the appearance of micropitting between driver versus driven gearing. The 

pinions saw more than twice the number of cycles as the gears, therefore the effects were more prominent.
67

Figure 7.7. Involute gearing showing early micropitting; both parts (circled) drive their respective mates

Only a summary of the surface fatigue failures appears here, a detailed discussion appears in Appendix A. 

After 86 hours of operation (Test 3) a significant amount of vibration was noted, prompting shutdown of 

the test fixture. Upon inspection, pinion 1 in gearbox 2 was found to have significant macropitting (see 

Figure 7.8 through Figure 7.10). The origin of the pitting appears to fall at the pitch line of the tooth flank, 

which is consistent with the area of highest calculated stresses for involute gearing. This pinion was 

operated as a reducer—the pinion drives the gear. Note that there is frosting, or micropitting, in the 

dedendum (near the tooth root) of the tooth flank. In some cases, this phenomenon has been thought to 

precipitate a macropitting failure, at the pitch line where the macropitting originated no micropitting is 

present. The macropitting therefore can be considered a classic fatigue failure of the type expected in this 

test. It is important to note that only 10 hours of operation occurred between inspections. There was no 

hint of damage prior to the high vibration level (manually monitored at that time) that prompted cessation 

of the test.

Figure 7.8. Massive macropitting on pinion tooth flank (Test 3, gearbox 2, pinion 1)
68

Figure 7.9. Additional pitting

Figure 7.10. Further pitting and opposite flank scuffing damage from Test 2

The definition of macropitting failure for this test was set as 1.5% of the operational tooth. In this failure 

the amount of pitting greatly exceeded the definition. It was expected that macropitting would progress 

from small pits to larger ones and that, with the increase in pitted area, the measured vibrations would 

increase. At the time of this failure, however, the test parameters and warnings had not yet been set up. 

Test 4 was terminated after 18 hours due to another surface fatigue failure. Figure 7.11 and Figure 7.12 

show the failure—again, a pinion operated as a speed reducer, this time in gearbox 1. Here the surface 

damage was found during a routine inspection. The amount of damage is indicative of the definition of 

failure. The total test time for this pinion was 104 hours, or 13.7 million cycles. 

It is important to note the extensive micropitting on the tooth flanks. For these parts, it occurs almost 

exclusively in the dedendum of the gear tooth (below the pitch diameter). There has been much discussion 

in the gear industry concerning whether the type of action (speed increaser versus speed reducer) affects 

the appearance of micropitting, or whether it is entirely a result of load, lubrication, and surface finish. 
69

During testing, it appears that there is a difference due to the type of action. Both of the failed pinions 

were loaded as speed reducers (for example, the pinion drives the gear) and there was extensive 

micropitting in the dedendum of the tooth flanks. The other pinion, which was located in the same 

gearbox, operated for the same number of load cycles as the failed pinion in Test 4, and was loaded as a 

speed increaser (Figure 7.13), shows much less wear on the active flank.

Within the gear industry, there is much discussion on the affects of gear geometry, how it is loaded 

(increaser or reducer), and wear potential. A preliminary analysis of the probability for wear on the 

involute gear set (first with the pinion driving the gear; and second with the gear driving the pinion) 

revealed identical specific film thicknesses and similar low probabilities for wear. Outside of the edges 

(tip/root) of tooth contact, the calculated film thicknesses are the same for both analyses. The film 

thickness is a function of the relative curvatures between contacting surfaces, therefore the direction of 

rotation does not appear to have an effect. Additional study into this area is required before definitive 

conclusions can be drawn. 

Figure 7.11. Failed involute pinion from Test 4

Figure 7.12. Pinion loaded as a speed reducer 
70

Figure 7.13. The other pinion in the same gearbox loaded as a speed increaser and showing minimal wear

It is thought that micropitting might precipitate macropitting, therefore the presence of this wear is 

important. Contact loads on the surface of a gear tooth flank are directly related to the relative curvature 

between that flank and the mating flank. Standard involute gears have a convex shape, therefore the 

contact patch between the parts is relatively narrow. Further, the convex-convex contact of the gear teeth 

acts to “pump out” lubricating oil, reducing the effective film thickness. In contrast, mating Convoloid 

gear teeth have a convex-concave contact that greatly reduces the relative curvature, creating a wider 

contact patch under load which maintains a greater amount of oil between the surfaces. This increase in 

oil-film thickness is expected to reduce or eliminate micropitting in Convoloid gearing. Further, the 

Convoloid tooth form has no contact in the high-stress area of the pitch diameter. The design method 

allows Convoloid teeth to have a consistent contact stress across the entire working flank, which is in 

contrast to the varying stresses found in involute gear teeth. 

Another aspect that has generated interest is finding out if the life of a gear is affected by whether the gear 

component is the driving element (pinion for a speed reducer and gear for a speed increaser) or the driven 

element (gear for a speed reducer and pinion for a speed increaser). To this point, all failures have been of 

the driver element (pinions loaded as an increaser), and early small macropitting is visible on a driving 

gear loaded as an increaser.

A third macropitting failure occurred after a short period of testing. Once again, the failed member was a 

driving pinion—this time with 82 hours of operation. To date, the only macropitting failures have been 

found in driving pinions having 86, 104, and 82 hours of operation. The other involute pinions (being 

driven by the gear) have survived for more than 168 hours and had minimal visible wear. A small 

macropit is visible on a driver gear; however, this failure has not reached the test definition of failure. All 

remaining gear flanks show minimal wear.

7.2. Variations During Test

A few situations in which variability might have been introduced to the test must be noted. 

7.2.1. Incorrect Torque Application

Early application of load to the test fixture (necessary to confirm operation of the data gathering) caused a 

greater test load than intended (7,880 lb-in). No ill effects were noted, however, and testing continued. 
71

7.2.2. Loss of Lubrication

In two instances, lubrication of the test gearboxes was lost during operation. One loss affected only 

gearbox 2, damaging the Convoloid gearing. A second instance was caused by a loss of power to the 

entire test fixture, damaging all gearing. Although specific data from these gear sets is not available for 

life comparisons, the operational time (~160 hours) proved to be more significant after experiencing 

surface fatigue failures in parts having shorter life spans.

7.3. Test Results

A Convoloid test set was destroyed due to a test rig malfunction, therefore time remaining on the contract 

permitted testing of involute gears only. Although the accumulated life of each involute gear and pinion 

was logged, only three involute gears could be defined as “test failures.”(See Section 6.2.6.2 for a 

definition.) As noted at the end of Section 7.1.2 (above), the failures occurred in a reasonably tight time 

span (82 hr to 104 hr), a fact that lends credibility to the consistent accuracy of the gearing, the 

capabilities of the test rig itself, and many other positive factors. The added fact that all three failures were 

“driving pinions” is not only another “positive factor,” but future testing using the three-hole 

configuration might shed further light on the relative failure rates of driving pinions versus “driven 

pinions.” Different theories abound regarding the behavior of tooth surfaces with respect to the direction 

of subsurface shear direction and failure modes and times as a result.

This overall performance provides some assurance that continued testing would produce solid 

comparisons on the relative performance of involute gearing versus Convoloid (see Appendix A for 

details). Additionally, these three failures have provided detail on the life cycle–stress relationship 

through a spreadsheet developed by Genesis Partners. The life factors developed will help guide future 

testing protocols for the “three-hole continuation,” as well as that for general gear testing (see Appendix 

A for details). 

7.4. Continued Future Testing

Appendix A provides detail on the expected times and procedures to complete the test program so that 

involute and Convoloid performance can be compared with a reasonable degree of statistical surety. 

8. Detailed Design of Larger Gearboxes Using Convoloid Gear Sets

8.1. Selection Criteria

The method of selection for the most advantageous existing wind turbine gearbox to study and then to 

retrofit with gearing designed with the Convoloid tooth form was based on the following criteria. 

• Gear system architecture to be common to most gearboxes 1.0 MW and greater. 

• The existing substantive data on the market-sell price for the particular model selected. 

• The substantial detail on the existing gear designs enabling the proper assessment of existing 

gearing stress levels. 

• General design detail is available—such as housing sizes, enclosing configuration, existing 

bearings, carrier configuration, and shaft configurations, so that solid modeling can be well 

established; and

• Genesis Partners has substantive detail on existing bearing types and sizes of the selected model to 

help assess cost and bearing load/life calculations.
72

8.2. Extrapolation

The basic steps in application of the data developed in the Micon 108 test (not enough data has been 

collected from the 3-Hole Test as of the date if this writing) to allow reliable extrapolation to wind turbine 

gearboxes of larger sizes are as follows.

• As a result of early Convoloid test programs, the stress concentration factor protocols for 

Convoloid gearing were changed and then used in the Micon 108 test program with great success. 

• The calculation of surface durability stress levels in Convoloid gearing appear to be very 

reasonable. No macropitting was observed in executing the Micon 108 protocol. 

• Tip relief calculations as a result of the Micon 108 tests have been refined to compensate for the 

results of the test.

• Lead modification protocols using an involute-based LVR program worked moderately well for 

the 108 tests, with major improvements projected for the future. 

The gearbox that fit all of the above criteria was a 750-kW gearbox commonly used in the wind turbine 

industry today; accordingly it was selected for the study.

8.3. Comparison of Original 750-kW Gearbox with Convoloid Enhanced Gearbox

The first step in the redesign process was to study in depth the geometry of the gearing existing in the 

750-kW gearboxes. Detailed gear data was available including module, number of teeth, helix angles, and 

face widths. A solid model of the gearbox was developed (Figure 8.1) that corresponded with the weight 

of the physical part. 

Figure 8.1. Representative 750-kW gearbox

It was assumed that all materials conformed to AGMA 2001 Grade 2 carburizing steel, and that gear 

accuracies were very high. These two parameters were held constant throughout the analysis. 
73

The following data was reverse-engineered using AGMA/ISO procedures for all gears. 

• Bending stresses

• Surface durability stresses

• Radial, tangential, and thrust loads 

Input torque and speed to the low speed planetary stage were calculated at full load (750 kW) and 

standard input speed for a wind turbine gear drive of this power rating (28.6 rpm). From this information 

bearing loads could be calculated followed by the calculation of bearing L10 lives. 

It was an additional constraint on the redesign that the planetary ratio and (2) successive helical pair speed 

increasing stage ratios were distributed in the Convoloid gearbox about equal to those of the involute 

gearbox. The Convoloid gears then were designed using the same input torques and speeds, but reducing 

the general size of the Convoloid gears until surface durability stresses were at or below those of the 

involute gearbox at each gear set. Bending stresses were not the limiting parameter in the design. Loadsharing protocols for the planetary stage followed AGMA practice and were applied consistently for both 

the Convoloid redesign and the involute gearbox. 

The next step was to select and rate the bearing complement of the Convoloid gears. Although the twospeed increasing helical gear pairs were straightforward in the design and rating of their supporting 

bearing complements, the planetary carrier and carrier bearings were not. In this case, the resulting 

compact sun, planet, ring configuration of the redesigned Convoloid gears drove the bearing forces 

considerably higher than could be supported adequately using the involute configuration. The carrier and 

planet bearing complement had to be redesigned to achieve the desired bearing life. See Figure 8.2 for the 

solid model carrier rendition and Table 8.1 for a summary of the important parameter that resulted.

Table 8.1. Comparison Between an Existing 750-kW Planetary Stage and the Convoloid Design

Parameter Existing 750-kW Gearbox Convoloid Enhanced Gearbox

Planet Speed 34.12 rpm 34.12 rpm

Planet Torque 34.9 kN-M 34.9 kN-m 

Bearing Type Spherical Roller (2) Cylindrical Roller (4)

Bearing Configuration Under Planet Straddle Mount

Bearing Loads 367.2 kN 360.9 kN

Bearing L10 life 129,900 hr 124,900 hr

Planet Carrier Weight 775 kg 646 kg (-16.7%)

Planet/Brg Assembly 225 kg 218 kg (-3.1%)

Ring 503 kg 456 kg (-9.3%)

Next, the entire assembly was modeled and enclosed with a housing very similar to the reference 750-kW 

gearbox, but smaller. 
74

Figure 8.2. Redesigned low speed (planetary) using Convoloid tooth profiles (750 kW gearbox) 

8.4. Analytical Model Cost Savings and Extrapolated Results

The fully assembled 3-D Convoloid model was “weighed” electronically and compared to the original 

750-kW weight calculation. The resulting weight reduction was 34.9%. Knowing that additional factors 

must be applied to realistically estimate the actual “in practice” weight savings, a generous safety factor 

was applied bringing the projected weight savings for calculation of COE values to 21.0%. Weight of a 

sophisticated mechanical assembly (such as a gearbox) drives the price, therefore it is reasonable to 

assume that a price reduction of like power capacity, like configuration, and like manufacturing 

accuracies and materials would be 21%. 

9. Cost of Energy Analysis—Summary

9.1. Calculation of Achievable Reductions in Size and Cost of Convoloid Versus 

Involute Wind Turbine Generator

Section 8.3 described how achievable reductions in size and cost of an upsized Convoloid gearbox could 

be accomplished. Following the 21% “achievable” decrease in weight of the Convoloid 750-kW gearbox 

compared to its established in-service involute counterpart, Genesis Partners developed extensive cost 

models from NREL publications and Genesis’ own market research, including actual quotation figures for 

reasonable production quantities of this specific gearbox. With this information in hand, Genesis Partners 

made what it thinks are reasonable upscaling assumptions. 

• The weight of a gearbox per megawatt is relatively constant up to the 3-MW to 5-MW range. 

• Weight is directly proportional to cost.
75

Further, a conservative cost factor was applied to the already factored weight savings (providing the 

“achievable” 21% weight reduction) to allow for pricing contingencies. The resulting cost savings with 

Convoloid designs was calculated out at $35,420/MW (2008 dollars) of capacity, compared to its involute 

equivalent as a baseline.

9.2. Calculations of Advantages of Retrofitting Convoloid Gearing into Existing Involute 

Wind Turbine Generators

Section 3.3.3 projects O&M costs as a function of stress level. For there to be economic advantages of 

retrofitting existing involute gearboxes with Convoloid gearing, it follows that the slightly increased cost 

of Convoloid gearing—and most probably its bearing complement (due to the effort to balance the lives 

of the gears and the bearings)—must be less than the resulting O&M projected costs over a reasonably 

long lifetime of the wind turbine.

• Although of the two stress summaries of the Micon 108 high speed and low speed pass gears show 

torque increases of 46% and 33% for Convoloid designs (see Table 3.1, Table 3.2), to be 

conservative only a 25% increase in torque is considered for this example.

• For the existing power, the surface durability stresses are reduced by 11.8% with the bending 

stress reduction being proportional to the torque reduction at 25%. 

• This increase in bending strength does not limit the system, therefore the surface durability 

stresses drive the reliability and, thus, the O&M costs. This conclusion is dependent on the bearing 

ratings and other component designs keeping pace with the gear-life projections.

• A reduction in stress is directly related to an exponential increase in the anticipated lives of the 

gearing. This increase in life directly affects O&M costs due to not needing to replace the gearbox 

as often.

• According to Figure 2-17, page 57 of DOE Publication, “20% Wind Energy by 2030,” frequency 

of repairs to wind turbines of 2.5 kW to 1.5 MW totaled 2% for “drive trains” and 5% for 

“gearbox.” 

9.3. Cost of Energy Trade-Offs

In the trade-off calculations that follow, reference was made to the following documents. 

• Development of an Operations and Maintenance Cost Model. NREL Subcontract No. YAM-4-

33200-07 

• “Wind Energy Handbook,” Burton & Sharpe (2001) 

• “20% Wind Energy by 2030,” published by the U.S. Department of Energy Office of Scientific 

and Technical Information 

• “Wind Turbine Design Cost and Scaling Model” NREL/TP-500-40566, December 2006 

• “Wind Pact Turbine Rotor Design Study,” NREL/SR-500-32495, June 2002 

• COE Projection Sheet ($ 2002)—NREL

Three specific trade-offs were examined and analyzed for their respective COE implications. Each was 

put through the accepted wind energy calculations using the NREL LWST Baseline (2002) Turbine 
76

Calculations—Class 4. Each then was put through the standard calculation for COE using the NREL 

calculation sheet for the Baseline 1.5-MW Turbine. The results then were summarized and conclusions of 

the study were articulated. The Baseline COE Projection against which each trade-off was compared is 

shown in accompanying figures. For a complete analysis of COE calculations for various sizes of wind 

turbines, see Appendix B. 

9.3.1. Wind Energy Generation—Trade-Off 1: Gearbox Cost Reduction / Tower Height 

Increase (Near Net Cost Neutral)

Trade-off 1 was a study of trading off the lower weight of the Convoloid gearbox compared to its involute 

counterpart against a 10-m increase of hub height from 65 m to 75 m with an assumed net cost neutral 

result (see Figure 9.3, Figure 9.4). Using “Wind Turbine Design Cost and Scaling Model” and COE 

Projection Sheet ($ 2002) for reference, the Baseline tower cost is listed as $101,000. An authoritative 

source (the president of a major wind turbine manufacturer), indicated that the cost of the tower is 

proportional to the fourth power of any height increase. Thus, the estimated increase in cost to raise the 

Baseline tower from 65 m to 75 m is as follows.

4

75 $101, 000 $133, 099 65

  × =     (2002 dollars) 

The increase from $101,000 to $133,099 ($32,099) compares favorably with the decrease in estimated 

cost of the gearbox at $31,000 (Figure 9.3), and justifies an approximate trade-off because both figures 

are provided in 2002 dollars. 

Figure 9.1. Baseline wind energy with 65-m high tower equals 4,439 MWh per year; wind energy generation 

trade-off 1, tower height increase (COE benefit of 3.5%) 
77

Figure 9.2. Convoloid General Electric wind energy with 75-m high tower equals 4,598 MWh per year; wind 

energy generation trade-off 1, tower height increase (COE benefit of 3.5%) 
78

Figure 9.3. Cost of energy comparison for trade-off 1, tower height increase

9.3.2. Wind Energy Generation—Trade-Off 2: Gearbox Cost Reduction / Blade Diameter 

Increase (Near Net Cost Neutral)

In this analysis, the hub height was held fixed at 65 meters, but the blade diameter was increased to 

75 meters from 70 meters with a cost neutral result (see Figure 9.5, Figure 9.6). Using reference numbers 

(4), (5), and (6) above, per-blade cost increases are calculated as follows.

For the 75-meter diameter blades:

Material: .4019R3—955.24 = $20,238.81 

 Labor: 2.7445R2.5025 = $23,849.34 

 Subtotal T = $44,088.05 

 Total per blade = T ÷ 0.72 = $61,233.41 

For the 70-meter diameter blades:

Material = $16,276.22 

 Labor = $20,067.51 

 Subtotal T = $36,343.73 

 Total per blade = T ÷ 0.72 = $50,477.40 

Difference per blade = $61,233.41 (75m) 

 - $50,477.40 (70m) 

 $10,756.01 

Total for (3) blades = $32,268.00 
79

This figure compares favorably with the decrease in estimated cost of the gearbox at $31,000 (Figure 9.6), 

and justifies an approximate trade-off because both figures are given in 2002 dollars. 

Figure 9.4. Baseline wind energy with 70-m diameter blades is 4,439 MWh per year; wind energy trade-off 2, 

blade diameter increase (COE benefit of 7.5%) 

Figure 9.5. Convoloid GB Wind Energy with 75-m diameter blades is 4,798 MWh per year; wind energy 

generation trade-off 2. blade diameter increase (COE benefit of 7.5%) 
80

Figure 9.6. Cost of energy comparison for trade-off 2—blade diameter increase
81

9.3.3. Wind Energy Generation—Trade-Off 3: Gearbox Cost Reduction / Blade Diameter 

Increase (Near Net Cost Neutral)—Reduced Wind Speed

In this analysis, the hub height was held fixed at 65 meters but the blade diameter was increased to 

75 meters from 70 meters with a cost-neutral result. Further, the nominal wind speed was reduced until 

the energy capture equaled that of the Baseline 1.5-MW unit at 4,439 MWh per year (Figure 9.7).

Figure 9.7. Baseline with 70-m diameter blades W.E. with Class 4 wind speed (5.8 m/s) equals 4,439 MWh/yr

Figure 9.8. Convoloid GB with 75-m diameter blades W.E. with low wind speed (5.539 m/s) equals

4,439 MWh/yr
82

9.4. Best Trade-Off Model

The effects of these trade-offs are summarized in Table 9.1. 

Table 9.1. Summary of Wind Energy Generation Options

Turbine Configuration COE Cost Reduction

Baseline configuration—5.8 m/s wind speed (65-m tower / 70-m 

diameter rotor) 0.0480 — 

Trade-off 1—5.8 m/s wind speed gearbox savings / increase tower 

height (75-m tower / 70-m diameter rotor) 0.0464 3.5%

Trade-off 2—5.8 m/s wind speed gearbox savings / increase rotor 

diameter (65-m tower / 75-m diameter rotor) 0.0444 7.5%

Trade-off 2—slower 5.539 m/s wind speed gearbox savings / 

increase rotor diameter (65-m tower / 75-m diameter rotor) 0.048 — 

From this data, two distinct advantages can be highlighted (tabulated in Table 9.1). The first is that, for 

best performance, trade-off 2 is decidedly the best, offering a cost reduction in COE of 7.5%. The second 

advantage is that which most explicitly meets the objectives of the NREL LWST program, that is, to 

produce cost-effective energy capture at lower wind speeds. Using alternative 3, an energy capture of 

4,439 MWh per year can be obtained with a conventional Baseline 1.5-MW turbine at 5.8 m/s wind speed 

(Class 4) (70-m diameter rotor) with a conventional gearbox or 4,439 MWh per year energy capture with 

a Convoloid gearbox incorporated into the design at 5.39 m/s (Class 3) (75-m diameter rotor).

9.5. Objectives of NREL LWST Development Program

The U.S. Department of Energy has a program goal to develop competitive wind power generation in 

regions of lower wind speeds—specifically Class 4 winds such as those commonly found in the 

Midwestern and western United States. Section 1.1.1 provides further detail. The energy capture of the 

cost-neutral Convoloid gearbox–blade diameter increase typified by alternative 3 using Class 3 winds 

(5.539 m/s) equals the energy capture of the classic involute Baseline design using Class 4 winds 

(5.8 m/s). This result exactly matches original NREL LWST program objectives as directed to improve 

energy capture of the wind resources found in the Midwestern United States. 

9.6. Potential Effects of Convoloid Gearing Technology

9.6.1. Performance

Several alternatives exist to improve the performance of wind turbine generators (WTG) through the 

application of Convoloid gearing technology. At one end of the possible spectrum is the direct replacement of Convoloid gears throughout the gearing system of an existing wind turbine gearbox. In this case, 

the gear operating stresses both for surface durability and tooth bending are considerably lower than those 

of the comparable involute gearbox. Under like operating conditions, this should significantly reduce 

O&M costs due to the greatly increased life of the Convoloid gearbox (see Section 3.3.3 for details). 

There are additional extenuating conditions that must be considered to achieve full advantage of such 

Convoloid gear retrofits. 

• The bearing compliment stresses likewise must be examined and possibly redesigned to achieve a 

life expectancy consistent with that of the Convoloid gearing. If this analysis is not done, then the 

bearing compliment as originally designed could become the limiting factor in any O&M cost and 

life projections.

• Comprehensive lubrication analysis. 
83

• Analysis of other components conducted to ensure that no modifications due to the Convoloid 

design have negatively impacted these components. 

At the other end of the spectrum is a completely redesigned WTG, where the size of the Convoloid 

gearing and the respective operating center distances have been reduced to achieve bending and surface 

durability stresses at or below those of the involute gearbox it replaces. At the designer’s choice, new 

mounting interface provisions can be used or the existing mounting configuration of the involute gearbox 

being replaced can be maintained for interchangeability. Here, again, bearing design and ratings are 

extremely important considerations, as are other factors such as lubrication.

9.6.2. Economic Factors

The economic impact of Convoloid gearing technology applied to wind turbine gearboxes can be directly 

related to the design and manufacturing alternatives related in the above section. Should a direct 

replacement of Convoloid gears in place of the involutes in a gearbox be analyzed, the following is 

projected. 

• In accordance with Section 3.6, there is an additional cost of bearings in the bearing complement 

to achieve an expected life of the bearings coincident or nearly so with the life of the replacement 

Convoloid gearing. 

• The production costs of the Convoloid gears will be slightly higher than the involute set they 

replace primarily due to single-flank grinding times required for Convoloid gearing at its present 

stage of development compared with double-flank grinding capabilities of the gear manufacturing 

state of the art for involutes. 

At the other end of the spectrum is the option to completely redesign the gearbox using Convoloid 

technology. In Section 8.1, it was established that a conservative estimate of the cost savings of this 

alternative is 21% of the nominal cost of an involute gearbox. Through its market research and 

information gathering from reliable industry sources, Genesis Partners has established a nominal cost 

savings of $35,420 per MW (2008 $) of gearbox rated capacity. 

Convoloid gearboxes can be designed to be significantly smaller, lighter, and less costly (for equal gear 

stress levels) than involute gearboxes, therefore compromises in design approaches to optimize lower 

overall capital costs and O&M costs are possible. Designing for lower Convoloid gearbox stresses 

improves the O&M costs, provided that bearing ratings, costs, and field performance keep pace with the 

improvements inherent in the gearing. 

9.6.3. Extended Service Life

The significant increases in anticipated gear life have immediate impacts on the O&M costs. Increased 

gear life causes greater intervals between repair and replacement of the gearboxes. Longer intervals 

between repairs means that there will be fewer repairs within the expected service life of a turbine—

greatly reducing O&M costs. With proper maintenance, achieving the full wind turbine service life is 

possible with gearboxes containing Convoloid gearing. 
84

Appendix A. Three-Hole Test

This appendix provides supplementary design, rating, and performance data developed up to the cessation 

of the 3-Hole Test. Details of test instrumentation parameter are listed as are general features of the 

3-Hole Test rig. Descriptions of the actual test protocols, gear performance, and failure analysis are 

contained at the end of this appendix. 

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Involute Gear Rating per AGMA 2001

American Gear Manufacturers Association Time: 15:40:34 Date: 2008/11/20 

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*** ECHO OF INPUT ***

1 1 Data set number

2 Genesis 3-Hole Test

3 7NDP 16/35

4 

5 2 2 In/Out = 1, si = 2, en 3 both units

6 2 = 1 Short = 2 Long Output Form

7 1 = 1 External = 2 Internal Gear

8 1 = 1 Single = 2 Double Helical

9 20.000000 deg Normal pressure angle

10 7.0000000 in^-1 Normal diametral pitch

11 18.697400 deg Helix angle

12 4.0000000 in Working center distance

13 16 35 Pinion / Gear number of teeth

14 2.7930000 5.7570000 in Pinion / Gear tip diameter

15 1.5000000 1.4500000 in net face width (gap not included)

16 11 Transmission Accuracy

17 0.0000000 0.0000000 in Pitch variation

18 0.45865739 0.75120803 addendum modification coefficient

19 0.16000000 E-01 0.16000000 E-01 Tooth thinning for backlash (norm)

20 10000 10000 Number of teeth on tool

21 1.4000000 1.4000000 tool addendum (normalized)

22 0.82146946 E-14 0.82146946 E-14 tool addendum modification coeff

23 0.0000000 0.0000000 tool protuberance (normalized)

24 0.39360256 0.39360256 tool tip radius (normalized)

25 0.0000000 0.0000000 Stock allowance per side (norm)

26 32.000000 32.000000 mu in surface finish (Ra)

27 0.0000000 0.0000000 in rim thickness (0 = solid disk)

28 3.0000000 in Pinion bearing span

29 0.85000000 in Pinion offset from center bearing span

30 2200.0000 rpm Pinion speed (rpm)
85

31 7000.0000 lbs in Pinion torque

32 300.00000 hours Design life

33 1 1 number of mesh contacts

34 2 2 Idler use = 1 yes = 2 no

35 4 4 Surface hardness, = 1 HV = 2 HBN. . .

36 58.000000 58.000000 Surface hardness values

37 4 4 Core hardness, = 1 HV = 2 HBN...

38 30.000000 30.000000 Core hardness values

39 1 1 mat, = 1 steel, = 2 cast iron, ...

40 1 1 Material sub classes

41 3 3 Heat treatment = 1 TH/NA = 2 Flam..

42 1 1 = 1 quench & temper = 2 Normalized

43 0 0 pattern type =0 na =1 A =2 B

44 2 2 material grade

45 0.0000000 0.0000000 psi Allow. contact stress number

46 0.0000000 0.0000000 psi Allow. bending stress number

47 0.29000000 0.29000000 Poissons ratio

48 30000000. 30000000. psi Modulus of elasticity

49 2 Lapped adjusted assy = 1 yes = 2 no

50 1 Lead crowned/corrected = 1 yes = 2 no

51 3 = 1 open = 2 commer = 3 prec = 4 xtra

52 2 Point stress calc = 1 hpstc = 2 tip

53 0.0000000 Pitting resistance service factor

54 0.0000000 Bending strength service factor

55 1.0000000 Pitting resistance safety factor

56 1.0000000 Bending strength safety factor

57 1.0000000 1.0000000 Reliability factor

58 1.0000000 Overload factor (clause 9)

59 0.0000000 lbs Maximum tangential load

60 0.0000000 Load distribution factor

61 0.0000000 Load Distribution factor at yield

62 2 Application = 1 conser = 2 Indus

63 0.0000000 Size factor (clause 20)

64 0.0000000 Temperature factor (clause 19)

65 0.0000000 Surface condition factor

66 0.0000000 Dynamic factor (clause 8)

67 0.0000000 0.0000000 Life factor pitting

68 0.0000000 0.0000000 Life factor bending

69 0.85000000 0.85000000 Life factor pitting @ 10^10

70 0.90000000 0.90000000 Life factor bending @ 10^10

Rating Routine Error Messages See User’s Manual for More Information

1) WARNING: Calculated pitting safety factor for pinion ( 0.937) is less than input value. 

2) WARNING: Calculated pitting safety factor for gear ( 0.954) is less than input value. 
86

7) WARNING: Contact Load Factor (1959.1) exceeds allowable contact load factor for the pinion. 

8) WARNING: Contact Load Factor (2032.6) exceeds allowable contact load factor for the gear.

Effective Case Error Messages See User’s Manual for More Information

5) NOTE: Minimum case depth is calculated using contact stress above the maximum value 

recommended in the standard. 

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AMERICAN GEAR MANUFACTURERS ASSOCIATION Time: 15:40:34 Date: 2008/11/20

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SINGLE HELICAL PINION GEAR

** GEAR GEOMETRY **

Number of Teeth (-) 16 35

Normal Diameter Pitch (Module) 1/in (mm) 7.0000 (3.6286)

Normal Pressure Angle deg 20.0000

Helix Angle deg 18.6974

Op. Center Distance in 4.0000

Outside Diameter in 2.7930 5.7570

Face Width in 1.5000 1.4500

Effective Face Width in 1.4500

Gear Ratio (-) 2.1875

** GEAR GEOMETRY—NORMALIZED **

Addendum Mod. Coeff (-) 0.4587 0.7512

Tooth Thinned for B/L (-) 0.0160 0.0160

Stock Allow./Tooth Flank (-) 0.0000 0.0000

** TOOL GEOMETRY—NORMALIZED **

Add. Mod. Coeff of Tool (-) 0.0000 0.0000

Protuberance of Tool (-) 0.0000 0.0000

Addendum of Tool (-) 1.4000 1.4000

Tool Tip Radius (-) 0.3936 0.3936

Number of Teeth on Tool (-) 10000 10000

** LOADING DATA **

Design Life hours 300

Pinion Torque (input) lb in 7000.0000

Speed rpm 2200.00 1005.71

Pitch Line Velocity ft/min 1445.5433

Max Tang. Load (input) lb 0.0000

Type of Practice INDUSTRIAL

Type of Service PRECISION ENCLOSED GEARING

Reliability Factor(input) (-) 1.0000 1.0000
87

** ADDITIONAL INPUTS **

Type of Gear Set EXTERNAL

Bearing Span in 3.0000

Distance "s1" in 0.8500

Rim Thickness in 0.0000 0.0000

Dynamic factor (-) N/A

Transmission Accuracy No (-) 11

Abs. Pitch Variation (-) N/A

Lead Correction or Crown YES

Lapped or Adjusted NO

Number of Contacts (-) 1 1

Idler NO NO

Spur Loading N/A

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AMERICAN GEAR MANUFACTURERS ASSOCIATION Time: 15:40:34 Date: 2008/11/20

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SINGLE HELICAL PINION GEAR

** INPUT FACTORS **

Load Distribution Factor (-) 0.0000

Dynamic Factor (-) 0.0000

Surface Condition Factor (-) 0.0000

Overload Factor-Default (-) 1.0000

Size Factor (-) 0.0000

Temperature Factor (-) 0.0000

Load Dist. Ftr Overload (-) 0.0000

Pitting Stress Cycle Fac (-) 0.0000 0.0000

Bending Stress Cycle Fac (-) 0.0000 0.0000

Pitt Stress Cycle @10^10 (-) 0.8500 0.8500

Bend Stress Cycle @10^10 (-) 0.9000 0.9000

** MATERIAL DATA **

Material STEEL STEEL

Material Type PLAIN STEEL PLAIN STEEL

Material Grade or Class GRADE 2 GRADE 2

Heat Treatment CARB & HARDENED CARB & HARDENED

Induction Hard. Pattern N/A N/A

Quench QUENCH & TEMPER QUENCH & TEMPER

Surface Finish mu in 32.0000 32.0000

Modulus of Elasticity psi 30000000. 30000000.

Poisson’s Ratio (-) 0.2900 0.2900

Allow. Cont. Stress (input) psi 0. 0.

Allow. Bend. Stress (input) psi 0. 0.

Core Hardness Number (-) 30.00 30.00

Core Hardness Scale ROCKWELL C ROCKWELL C

Surface Hardness Number (-) 58.00 58.00
88

** MATERIAL DATA **

Surface Hardness Scale ROCKWELL C ROCKWELL C

Brinell Core Hardness (-) 287.16 287.16

Brinell Surface Hardness (-) 620.29 620.29

Allow Contact Str. No psi 225000. 225000.

Allow Bending Str. No psi 65000. 65000.

Elastic Coeff (lb/in^2)^0.5 2283.2142

** EFFECTIVE CASE DATA **

Process Fact / Core Coef psi 6400000. 6400000.

Maximum Effective Case in 0.0511 0.0508

Min. Eff./ Total Case in 0.0290 0.0290

Heavy Minimum in 0.0297 0.0297

Normal Minimum in 0.0225 0.0225

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SINGLE HELICAL PINION GEAR

** PRESSURE ANGLES **

Inv. of Norm.Press Angle (-) 0.014904

Std. Trans. Pressure Angle deg 21.0192

Normal Op. Press Angle deg 24.8704

Trans. Op. Press Angle deg 26.1712

Inv. of Trans. Press Angle (-) 0.017394

Trans. Tip Press Angle deg 36.2466 31.1421

Inv. of T. Tip Press Angle (-) 0.100516 0.060709

** PITCHES **

Transverse Dia. Pitch 1/in 6.6306

Normal Base Pitch in 0.4217

Transverse Base Pitch in 0.4423

Axial Pitch in 1.4000

** HELIX ANGLES **

Op. Helix Angle deg 19.3920

Base Helix Angle deg 17.5319

** DIAMETERS **

Ref. Pitch Diameter in 2.4131 5.2786

Operating Pitch Diameter in 2.5098 5.4902

Root Diameter in 2.1378 5.0868

Root Diameter (ref. 908) in 2.1378 5.0868

Base Diameter in 2.2525 4.9273
89

** TOOTH GEOMETRY **

Tooth Whole Depth in 0.3276 0.3351

Norm Tooth Thk (ref dia) in 0.2698 0.3002

Norm Top Land Thickness in 0.0913 0.0908

** CONTACT RATIOS **

Trans Contact Ratio (-) 1.2439

Axial Contact Ratio (-) 1.0357

Total Contact Ratio (-) 2.2796

** MISCELLANEOUS FACTORS **

Gear Ratio Factor (-) 0.6863

Effective Protuberance (-) 0.0000 0.0000

Minimum Contact Length in 1.8787

Fractional Part of m f (-) 0.0357

Adjusted No of Teeth (-) 16.8914 36.9500

Fractional Part of m_p (-) 0.2439

** LINE OF ACTION DATA **

Points Along LOA A B C D E F 

Dist (c_1-c_6) in 0.2756 0.3834 0.5535 0.7178 0.8257 1.7642

(gamma_a-e) -0.5021 -0.3073 0.0000 0.2970 0.4918

Rol. Ang (eqs)deg 14.0189 19.5057 28.1573 36.5189 42.0057

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SINGLE HELICAL PINION GEAR 

** MESH FACTORS **

Mesh Align Correction Factor (-) 1.0000

Surface Condition Factor (-) 1.0000

Mesh Alignment Factor (-) 0.0859

Lead Correction Factor (-) 0.8000

Pinion Proportion Factor (-) 0.0384

Pinion Proportion Modif (-) 1.1000

Load Distribution Factor (-) 1.1025

** AGMA 908 DATA (normalized) **

I-Factor (-) 0.1776

J-Factor (-) 0.5112 0.5456

Working Center Distance (-) 28.0000

Number of Teeth (-) 16.0000 35.0000

Effective Face Width (-) 10.1500

X-Factor (-) 0.4587 0.7512

Number of Teeth (tool) (-) 10000.0000 10000.0000
90

** AGMA 908 DATA (normalized) **

X-Factor (tool) (-) 0.0000 0.0000

Tool Addendum (-) 1.4000 1.4000

Tool Tip Radius (-) 0.3936 0.3936

Effective Protuberance (-) 0.0000 0.0000

Tooth Thinning for B/L (-) 0.0160 0.0160

Strength Conc. Factor (-) 1.5173 1.5956

** STRESS FACTORS SUMMARY **

Load Dist Factor-Ovrload (-) 1.0909

Number of Stress Cycles (-) 0.39600E+08 0.18103E+08

Reliability Factor (-) 1.0000 1.0000

Overload Factor (-) 1.0000

Hardness Ratio Factor (-) 1.0000 1.0000

Size Factor (-) 1.0000

Temperature Factor (-) 1.0000

Yield Strength Factor (-) 0.7500 0.7500

Stress due to Wmax psi 47827.85 42617.50

Allow. Yield Strength psi 105612.70 105612.70

Max. Tangential Load lb 5578.1250

Pitting Stress Cycle Fac (-) 0.9681 0.9861

Bending Stress Cycle Fac (-) 0.9931 1.0070

Dynamic Factor (-) 1.0903

Transmission Accuracy No (-) 11

Abs. Value of Pitch Var. in N/A

Calculated Driver Power hp 244.4444

Member Torque lb in 7000.00 15312.50

Max. Pitch Line Vel ft/min 10000.0000

Tangential Load lb 5578.1250

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SINGLE HELICAL PINION GEAR

************************** 

*** MAIN RATING VALUES ***

************************** 

** PITTING **

Allowable Transmitted Power at Unity Service Factor hp 214.4104 222.4548

Allowable Power at input Service Factor hp N/A N/A

Service Factor (input) (-) 0.0000

Service Factor (calc) (-) N/A N/A

Contact Load Factor psi 2233.4826

Allow Contact Load Fact psi 1959.0620 2032.5634

Contact Stress Number psi 232588.2285
91

** PITTING **

Allowable Power at input Safety Factor hp 214.4104 222.4548

Safety Factor (input) (-) 1.0000

Safety Factor (calc) (-) 0.94 0.95

** BENDING **

Allowable Transmitted Power at Unity Service Factor hp 263.0853 284.7086

Allowable Power at input Service Factor hp N/A N/A

Service Factor (input) (-) 0.0000

Service Factor (calc) (-) N/A N/A

Unit Load psi 26928.8793

Allowable Unit Load psi 28982.4234 31364.5225

Bending Stress Number psi 59975.8751 56197.9283

Allowable Power at input Safety Factor hp 263.0853 284.7086

Safety Factor (input) (-) 1.0000

Safety Factor (calc) (-) 1.08 1.16

** Power Summary **

Input Power hp 244.4444

Allowable Transmitted Power for input Service Factor hp N/A

Allowable Transmitted Power for input Safety Factor hp 214.4104

3-Hole Test Instrumentation Data Points

3-Hole Test Instrumentation Parameters

Notes

1. Involute gearbox is gearbox 1 (nearest the motor).

2. Convoloid gearbox is gearbox 2.

3. Shaft notation with respect to torque actuator or Lebow Transducer.

Recorded 

in Data 

File Alarms Signal Type

Convoloid Oil Flow Y Pulses

Convoloid Oil Temp Out Y 

Convoloid Oil Temp In Y

Convoloid Particles Y Excessive Particles

Convoloid Oil Pressure at Manifold Y Loss of Oil Pressure DC Volts

Convoloid, Outside, Lebow Bearing Temperature Y Excessive Temperature TC Isolated

Convoloid, Outside, Intermediate Bearing 

Temperature Y Excessive Temperature TC Isolated

Convoloid, Outside, Torque Actuator Bearing 

Temperature Y Excessive Temperature TC Isolated

Convoloid, Inside, Lebow Bearing Temperature Y Excessive Temperature TC Isolated

Convoloid, Inside, Intermediate Bearing Temperature Y Excessive Temperature TC Isolated

Convoloid, Inside, Torque Actuator Bearing 

Temperature Y Excessive Temperature TC Isolated

Convoloid Reservoir Temperature TC

Convoloid Magnetic Chip Detector Y Particles Detected DC Volts

Convoloid Filter Pressure Drop Y Excessive Filter Pressure Drop

Convoloid X Vibration Y Excessive Vibration AC mV
92

Recorded 

in Data 

File Alarms Signal Type

Convoloid Y Vibration Y Excessive Vibration AC mV

Convoloid Z Vibration Y Excessive Vibration AC mV

Convoloid Total Vibration (calculated) Y Excessive Vibration calc

Convoloid Heat Exchanger Oil Temp—Out TC

Convoloid Particle Count 4u DC Volts

Convoloid Particle Count 6u High Count DC Volts

Convoloid Particle Count 14u High Count DC Volts

High Speed Shaft RPM (Motor RPM) Y Pulses

Circulating Torque Y SG

Windup Force HP Y

Input Wattage Y

Ambient Temperature (Computer Side) Y TC

Ambient Temperature (Test Fixture Side) Y TC

Smoke Detector Y Smoke Detected

Setpoint Y

Bypass Valve Y

Hydraulic Actuator Pressure Low Pressure DC Volts

Involute Oil Flow Y

Involute Oil Temp Out Y

Involute Oil Temp In Y 

Involute Particles Y Excessive Particles

Involute Oil Pressure Y Loss of Oil Pressure DC Volts

Involute, Outside, Lebow Bearing Temperature Y Excessive Temperature TC Isolated

Involute, Outside, Intermediate Bearing Temperature Y Excessive Temperature TC Isolated

Involute, Outside, Torque Actuator Bearing 

Temperature Y Excessive Temperature TC Isolated

Involute, Inside, Lebow Bearing Temperature Y Excessive Temperature TC Isolated

Involute, Inside, Intermediate Bearing Temperature Y Excessive Temperature TC Isolated

Involute, Inside, Torque Actuator Bearing 

Temperature Y Excessive Temperature TC Isolated

Involute Reservoir Temperature Y TC

Involute Chip Detector Y Particles Detected DC Volts

Involute Filter Pressure Drop Y Excessive Filter Pressure Drop

Involute X Vibration Y Excessive Vibration AC mV

Involute Y Vibration Y Excessive Vibration AC mV

Involute Z Vibration Y Excessive Vibration AC mV

Involute Total Vibration Y Excessive Vibration calc

Involute Heat Exchanger Oil Temp Out TC

Involute Particle Count 4u DC Volts

Involute Particle Count 6u High Count DC Volts

Involute Particle Count 14u High Count DC Volts

Overview of 3-Hole Fixture Design and Shakedown

To free up manufacturing space, it was necessary to move testing outside of the original facility. Instead 

of locating the test in a fixed place, all equipment was fit into a 20-ft long shipping container (Figure 

A.1). Recently refurbished and customized with a man-door, window, heat exchanger openings (center 
93

bottom), and air-intake vents, this enclosure allowed the entire testing fixture to be set into place through 

the large end doors. All power and controls are enclosed and all that is necessary for testing is to connect 

power (460 v, 3-phase), air, and telephone/DSL.

Figure A.1. Test fixture enclosure

Manufacture of the test fixture was straightforward. The test fixture (Figure A.2) was covered with epoxybased paint inside. If testing must occur with synthetic oil, multiple fittings designed into the fixture allow 

ready access for instrumentation and oil flow. Not shown in the figure are the internal baffles or the 

lubricant deflection funnels used to direct oil falling out of the test gearboxes past the chip detectors. 

During assembly of the gearboxes to the test fixture, it was found that the mounting surface of the fixture 

had deformed when the top was welded to the tank. The resulting gaps were sufficient that oil would leak 

from between the gearbox and the surface. Additionally, potential harmful stresses could be applied to the 

gearbox housings (and subsequently to the bearings) if the mounting bolts are tightened beyond 

recommended torque values. Therefore, components were removed from the fixture and the mounting 

surface was milled. Subsequent assembly showed a much improved fit. All electrical relays and primary 

controls were contained in one enclosure (Figure A.2; Figure A.3). 
94

Figure A.2. The 3-Hole Test fixture after paint and initial top milling

Figure A.3. The 3-Hole Test electrical circuit assembly
95

Figure A.4. Electrical enclosure 

The computer control-system equipment was the same as used with the Micon 108 test, with a few 

additional control points—specifically, torque control and fire alarms sensors (see Figure A.5). After 

initial operation, temperatures were excessive around the computer equipment, requiring the installation 

of an air conditioner unit and separation of the test rig from the computers. The insulated barrier includes 

a polycarbonate window for visual monitoring of the test. 

Due to the long delay of achieving Convoloid gearing with good contact, both gearboxes were assembled 

with involute gearing and mounted to the fixture for shakedown and preliminary setup purposes. When 

the gearboxes were in place, the motor and other components could be installed. Extreme care was taken 

to align the various shafts, thereby reducing the risk of vibrations that could possibly affect data readings. 

After the shafts were aligned, the gearboxes and motor were pinned to the fixture with removable taper 

pins. When components needed to be removed during testing, they could be realigned easily. Removal of 

the gearboxes for part change outs (versus changing parts in place) is preferred, to keep out foreign 

contaminants and to be assured of proper bearing settings. 

After all the components were in place (Figure A.6), preliminary operation at modest speeds resulted in 

excessive vibrations that could be felt in the floor of the test container. The cause originally pointed to a 

bent shaft in the torque actuator. After repairs, however, significant vibrations still were evident. A 

number of couplings then were found to have severe run-out. Subsequent repairs—including adding 

machined alignment sleeves over the couplings—did not correct the vibration. 
96

Figure A.5. View inside the test container showing the computer monitoring system and polycarbonate 

view window

Figure A.6. Initial assembly of 3-Hole Test fixture

Further investigation revealed that the chosen couplings were not able to self-center due to the weight 

being carried between the two gearboxes (Lebow torque transducer on one shaft pair and the torque 

actuator between the other). Additionally, because both couplings on a given shaft (e.g., on either side of 

the Lebow) were flexible, a dynamically indeterminate load arrangement occurred causing the 

components in the middle to vibrate. (See Figure A.7.)
97

Figure A.7. The 3-Hole Test fixture showing couplings with alignment covers 

The vibration issue finally was solved by using different couplings with a rigid-flexible arrangement. This 

type of coupling reduces a degree of freedom, thereby better locating the coupled equipment (Figure A8, 

Figure A.9). Installation involves bolting the two halves together. With the bolt clearances, however, out–

of-balance vibrations again were felt. Careful installation using a dial indicator to center the bolted 

connections resulted in very smooth operation. 

Figure A.8. Fixed half of new coupling 
98

Figure A.9. Flexible half of new coupling

Installation of rigid-flexible type couplings in place of the previous coupling design resulted in greatly 

improved operation. No vibration from the actuator side was evident and light vibration was felt in the 

torque transducer side. Efforts to balance the system, however, were not fruitful. Investigation revealed 

that the internal bore was not perpendicular and was not centered in one of the coupling hubs, causing the 

transducer assembly to oscillate. To allow vibration monitoring (necessary for unattended operation) this 

coupling part required correction. Although the vibration precluded unattended operation, shakedown of 

the fixture continued. A certain amount of torque was desired to test the system instrumentation, therefore 

it was thought that a reasonable amount hydraulic pressure could be applied to the torque actuator 

(approximated to be 25% of the operational torque). After a period, some electronic verification was 

possible and it was determined that the actual load was nearer 45%.

As the test fixture warmed to operational temperatures (155°F to 160°F), pressures in the lubrication 

manifold of the Convoloid dropped well below (~20 psi) the specified pressure of 25 psi. (At the same 

time and temperatures, the involute pressures were only a slightly less than required.) Concerns regarding 

supplying lubricant to the gear meshes necessitated investigation of the lubricant supply pumps. These 

concerns are warranted because oil must be supplied to the gearing as they come into mesh. Depending on 

the load direction (see Figure A.10), the oil supplied to the gearing could come from the bottom. Without 

enough pressure there is the possibility of insufficient lubrication. 
99

Figure A.10. Cutaway of the 3-hole gearbox showing the gear mesh lubricating method

Positive displacement pumps require a bypass valve to prevent over-pressure during start-up with cold oil. 

At high pressures, lubricant is diverted from the high-pressure side of the pump back to the low-pressure 

side, causing a portion of the oil to circulate within the pump. As the pressure decreases the valve closes. 

The valve is controlled by spring pressure created by a set-screw. This screw is covered with a cap to 

prevent air leakage and contamination.

When the cap was being removed, the adjusting set-screw stuck to the cap and, inadvertently, it also was 

being removed. At some point, the pressure fell enough that no oil was being fed to the gear meshes. The 

fact that the set screw was being removed was not realized until it fell out of the pump housing—which 

caused 160°F oil to flow out of the machine. The machinery was shut down immediately and the pump 

was repaired. Inspection of the gear meshes, however, revealed extensive scuffing (welding and tearing of 

the loaded tooth flank) (Figure A.11; Figure A.12). 
100

Figure A.11. Severe scuffing on the Convoloid pinions due to loss of lubricant

Figure A.12. Severe scuffing on the Convoloid pinions due to loss of lubricant

These flanks no longer could be used for testing, therefore it was thought that continued operation for 

shakedown purposes would not further damage the parts. As expected, when under no load the Convoloid 

gearbox was louder than the involute gearbox. After applying the same load as used previously, however, 

the Convoloid gearbox became quieter than the involute gearbox. 

As instrumentation was calibrated, the torque-monitoring equipment revealed that the system was 

operating considerably above the expected test torque. The applied torque was reduced immediately, but 

the system had operated for approximately 10 hours at the greater load. The system operated for 5 more 

hours before an unexpected noise in the Convoloid gearbox caused the test to be shut down. Initial 

inspection of the inside of the gearbox revealed an interference pattern on the unloaded flanks. The 

gearbox was disassembled, and it was found that the intermediate shaft had broken. It is interesting to 

note that the flanks with severe scuffing appeared to have smoothed during the operation at lower torque 

levels.

An immediate investigation into how to prevent a similar incident led to (1) An increase in pump volume 

to maintain the requisite 25 psi at the distribution manifold; and (2) A redesign of the intermediate shaft to 
101

strengthen the bearing shoulder area. To address the low oil pressure, pumps with increased capacities 

were installed. Subsequent operation confirmed proper flow (Figure A.13). 

Figure A.13. The 3-Hole Test gear mesh oil nozzles with proper oil pressure (rotated to show flow)

Redesign of the intermediate shaft assembly commenced immediately following the failure. Safety factors 

of the new shaft design are nearly two times greater than with the original design. The new design 

required a trade-off with bearing size; however, with the controlled lubricant system and replacement of 

bearings planned with every gear change, the reduced bearing life was deemed to be sufficient for the test 

period. Manufacture of two shafts began immediately and four additional shafts were made to facilitate 

easier gearing changeover and longer shaft life. 

Inspection of the coupling hubs on either side of the Lebow torque transducer revealed that the bores in 

the solid hubs were not round, nor where they perpendicular to the mounting face. This caused the 

vibration that was experienced during operation and explains why the hubs were so difficult to install. 

Subsequent balancing of the hubs showed that minimal correction was needed. 

To allow the instrumentation to be qualified and to enable unattended operation, both gearboxes were 

assembled with involute gear sets. The number of cycles and load amount were logged for each loaded 

flank, so all test cycles could be tracked. Any gearing failures were documented to help generate a 

baseline for future Convoloid comparisons. A photo of the final test fixture configuration is shown in 

Figure A.14. 

Testing progressed with both gearboxes using involute gearing. Completing the programming of the 

control system was of prime importance. Additionally, until the safeguards were in place—including 

emergency shutdown and notifications—unattended operation is impossible. These safeguards are 

necessary to prevent the failure of instrumentation or test equipment from catastrophically affecting the 

test fixture.
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Figure A.14. Operational configuration of the 3-Hole Test fixture

The control system was designed to monitor critical aspects of the test fixture. Accelerometers in three 

planes are used to gather vibration data to determine any pending failures. (A view of the data screen for 

the vibration collection is shown in Figure A.15.) In previous testing, changes in the vibration level of the 

minimum amplitude planes correlated closely with bearing failures, and they also should give notice of 

pending gear wear. If wear begins causing loss of gear material, then chip detectors that are located in the 

oil flow from the gearboxes trigger an immediate shutdown of the test. Upon inspection of the gear teeth, 

additional testing can occur until the amount of tooth wear reaches 1.5% of the active tooth flank. 

Figure A.15. Sample of the monitoring software page (vibration shown)

To catch system failures before they lead to damage of the test gearing, numerous sensors with alarm 

settings have been included. Among these are bearing temperature, lubricant flow and pressure, lubricant 
103

temperature, and applied torque. Additionally, external sensors recently were installed to shut down the 

system if smoke or fire occurs within the container. Lastly, a fail-safe system has been added—the drive, 

torque, and lube systems will shut down if the computer controller becomes damaged or ceases to operate. 

When any failure mode occurs, notification is sent to the test supervisors and Genesis personnel so they 

can inspect and evaluate the alarm condition. 

One area of concern was found during the test operation: The ambient temperature within the test 

container exceeded 115°F after only 1 hour of having the test container closed. This occurred on a 65-

degree spring day. Having such high temperatures is excessive for the instrumentation electronics, thus 

necessitating the installation of an air conditioner for cooling and a barrier between the test fixture and the 

test computers. It was important to protect the monitoring equipment because its failure could cause the 

test fixture to “run away.” This led to the addition of the fail-safe circuit that is held closed as long as the 

computer is operational. If the electronic equipment fails the circuit opens, thus shutting down the system. 

Continued discussion of the test fixture operation is included below. 

Micropitting Test Protocol

3-Hole Gearbox Test

Low Wind Speed Turbine (LWST) Project Phase Ii 

Component Development 

Revision 1 

Modified 6-13-05 

NREL Contract #ZAM-5-33200-12 

April 9, 2007 

Initial testing in the 3-hole gearboxes is intended to qualify profile corrections in the Convoloid tooth 

form. This testing is in preparation for the main extended comparison test. Previous test gears developed 

micropitting early in the test cycle. This failure mode is not in itself catastrophic but it can lead to 

macropitting and, in extended operation, to complete failure of the gear. Before extended testing can 

occur the micropitting phenomena must be solved. This procedure applies only to the initial micropitting 

test. Subsequent testing is addressed in later procedures. This procedure assumes that all equipment and 

instrumentation has been assembled and tested in the static state (including a pressure test of the fixture). 

The gearboxes and oil distribution manifolds should not yet be installed. Gear teeth should be uniquely 

labeled prior to assembly into the gearboxes. 

First Start Procedure

1. Familiarize test personnel with instrumentation and test protocol. Set up and start manual test log. 

2. Obtain and label one 3-oz sample of new oil from drum for later testing.

3. Fill each tank to the appropriate level with test oil. Oil should be pumped out of the barrel through a 

3μ filter.

4. Activate the circulation pump to suspend any particles, and heat oil to 150°F. 

5. Channel the outlets from the filters back into the sump, and activate primary lubrication pumps. 

Allow to run for at least 4 hours, or until the oil cleanliness reaches -/15/12. 
104

6. Stop the lubrication pump (circulation motor should remain on) and attach lubrication manifolds and 

oil lines. Operate the pump to flush any particles then stop the pump.

7. Install the gearboxes, torque actuator, coupling, and motor. 

8. Assemble the lubricant manifold and complete piping. 

9. Ensuring no rotation of the gearboxes, start lubrication pump and verify lubricant flow through the 

housing. Check gearing spray-nozzle operation. Slowly hand-turn the shafts while lubricant flows, 

to assure no particles are present in the bearings.

10. Assure that all covers are in place and sealed.

Micropitting Test Procedure

1. Apply marking compound (bluing) to teeth flanks, and verify that each tooth is uniquely labeled. 

2. Assure that the circulating and lubricant pumps are operating. 

3. Take one 3-oz oil sample from each gearbox and perform a preliminary patch test.

4. Energize drive motor and slowly bring it up to operating speed (2,250 rpm). 

5. Verify that all components are running smooth with minimal vibration. 

6. Bring the load up to 10% of test torque (630 lb-in) and operate for 1 hour. Confirm that 

instrumentation operates properly. Continuously monitor particulates; if they exceed -/18/13, shut 

the system down and determine the reason. Increase load to 25% of test torque (1,575 lb-in) and 

operate for 1 hour. 

7. Stop operation and inspect wear pattern. Photograph the results. If any wear is visible through the 

transition zone, evaluate the extent of the micropitting. The test can be stopped at this point for 

further analysis and modifications to the profile correction. Some micropitting due to lead variations 

is acceptable (up to 10% of the active face).

8. Restart and slowly increase the load to 50% of test torque (3,150 lb-in) and operate for 1 hour. 

9. Stop operation and inspect and photograph wear pattern. 

10. Repeat steps (7) and (8) for 75% (4,725 lb-in) and 100% (6,300 lb-in) load with inspections and 

photographs after each test. Note any abnormalities or early wear. If any micropitting is visible, then 

continuation operation of the test must be evaluated. Further analysis and profile correction could be 

required. 

11. Once operation is proven at 100% of load, restart and operate for 24 hours and then re-inspect for 

micropitting.

12. If no micropitting is observed after the 24-hour test, then testing can progress to the extended 

comparison test.

Note that the bulk oil temperature should not exceed 155° F, measured at inlet to gearbox, at any time. 

Nor should any bearing temperature exceed 190° F (an alarm will sound). If either of these conditions 

exists, stop the test run and contact the instrumentation engineer. 

Inspection of gear teeth will require removal of the cover plate. Inspection lights should be used at various 

angles to adequately view the tooth flanks. Photos should be in focus and of high enough resolution that 

subsequent analysis is possible. If wear is visible, then a decimal inch scale should be included in the 
105

photo (for comparison). Scale markings should be readily visible as should the tooth wear. Be aware of 

flash glare on the tooth surface.

If there is a shutdown due to mechanical or lubrication problems, the First Start Protocol is repeated, 

starting with step 5. The micropitting test procedure, at the discretion of the test supervisor, then 

commences either from the beginning or from the last successful load step. If there is a shutdown due to 

instrumentation problems, the protocol is repeated from the last successful load step. (Repeat one load 

step.) 

Extended Comparison Test Protocol

3-Hole Gearbox Test

Low Wind Speed Turbine (LWST) Project Phase II

Component Development 

Revision 1 

Modified 6-13-05 

NREL Contract #ZAM-5-33200-12 

April 9, 2007 

The initial (shakedown) testing should be completed as specified in the Micropitting Test Protocol. Once 

complete, testing progresses to a long-term phase where a macropitting failure on one of the gear tooth 

flanks is anticipated. The intent of this procedure is to specify a method for extended testing with minimal 

physical supervision (unmanned), that still provides for proper control and inspection. The definition for a 

macropitting failure for use in this protocol follows guidelines specified in AGMA 925-A03 and is 

defined as damage (macropitting) occurring on a loaded tooth flank that is approximately 1.0-mm in 

diameter.

Full load test torque is 6,300 lb-in, however this torque can vary as needed to achieve a pitting failure in a 

reasonable test time. Test motor speed is 2,250 rpm. Variations in torque and speed are allowed as 

required for proper operation; however, all changes must be approved by Genesis Partners personnel. 

The point of reference for determining the direction of rotation of the test is from the fan end of the motor. 

This could be considered the “back” because it is not the end with the motor shaft, but this terminology 

has a different definition in the electric motor world. Figure A.16 shows the point of reference for 

determining direction of rotation.

Operation of this 3-Hole Test allows two different mesh conditions (a reduction gear set and an increaser 

gear set) in each gearbox (involute and Convoloid). Therefore, four gear meshes were tested at a time. 

Additionally, subsequent testing can be conducted by reversing the direction of rotation of the motor. 

Thus, sufficiently descriptive and unique annotations are required for each component as well as each 

tooth flank. Referencing Figure A.16, gearing component designations could follow the descriptions 

provided below. 
106

Involute Gearbox Convoloid Gearbox

1 Involute Pinion 1 5 Convoloid Pinion 1

2 Involute Gear 1 6 Convoloid Gear 1

3 Involute Pinion 2 7 Convoloid Pinion 2

4 Involute Gear 2 8 Convoloid Gear 2

Gear sets 5 to 6 and 3 to 4 act as increasers and gear sets 1 to 2 and 7 to 8 act as reducers. (For 

shakedown, see “Micropitting Test Protocol.”)

Figure A.16. 3-Hole Test fixture (top view)

Extended Comparison Test Procedure

1. Gear teeth (including each flank) should be uniquely labeled prior to assembly of the gearboxes. 

2. Ensure that the circulating and lubricant pumps are operating. 

3. Take one 3-oz oil sample from each oil reservoir and perform a preliminary patch test. Retain the 

samples for comparison purposes. 

4. Energize drive motor and slowly bring it up to operating speed (2,250 rpm). 

5. Verify that all components are running smoothly and with minimal vibration.

6. Bring the load up to 100% of test torque in 25% increments (1,575 lb-in, 3,150 lb-in, 4,725 lb-in, & 

6,300 lb-in). Confirm that instrumentation operates properly. Continuously monitor particulates; if 

they exceed -/18/13, then shut the system down and determine the reason. If particles are detected in 

the chip detector, then the test should be stopped. All components should be inspected and the cause 

for stoppage investigated. If no gear or bearing damage is found that would compromise the test, 

ensure that the circulated oil cleanliness is -/15/12 and restart the test.

7. An inspection should occur after approximately 165 hours (1 week) to verify tooth contact and wear 

(or lack thereof). The inspection should involve removing the coverplate and photographing the 

contact flanks of each gear/pinion. If any wear is evident, a metric (millimeter increments) scale 
107

should be included with the pictures. A 3-oz oil sample should be taken from each oil reservoir and 

retained for 2 months. Conduct a patch test and archive the results. 

8. Subsequent inspections should occur approximately every 335 hours (2 weeks) following the above 

procedure. When wear appears on a tooth flank, however, the inspection interval should be reduced. 

9. Once a member of one of the gear sets exhibits a macropitting failure, the test should be halted. An 

oil sample (with patch) should be taken and stored with the gear set. Notes for that specific gear set 

(loads, increaser/reducer, number of hours to failure) should be recorded. After inspection of all 

gearing, a decision on how to proceed must be made. In most instances two options exist (assuming 

a non-catastrophic tooth failure). 

A. The failed gear set should be removed from the test gearbox and be replaced by a new or 

undamaged gear set. Label and archive the gear set removed. To prevent variations due to 

possible bearing damage, new bearings should be used with the new gear set. Other wear items 

should be replaced as necessary. The remaining components in the test are assumed to still be 

in test, therefore great care should be exercised to prevent damage to these components. 

Additional testing on the removed component is possible (changing the motor direction of 

rotation, thus loading the opposite tooth flank), therefore removed components must be 

lubricated and stored carefully. 

B. Reverse the motor rotation to conduct testing on the opposite tooth flanks. Prior to restarting 

the test fixture, documentation should refer to the load and number of hours sustained by the 

previously loaded flank. If no wear is visible after subsequent testing, damaged components 

can be removed and be replaced with those with no visible wear. In this manner, additional 

cycles can be applied to various test components until tooth failure occurs. 

10. After the gearbox has been reassembled, circulate lubricant while the gearbox is at rest to remove 

any contaminants introduced during assembly. When the oil cleanliness reaches -/15/12, restart the 

procedure at step 3. 

Note that the bulk oil temperature should not exceed 155° F—measured at inlet to gearbox—at any time. 

Nor should any bearing temperature exceed 190° F (an alarm will sound). If either of these conditions 

exists, stop the test run and contact the instrumentation engineer. 

Any change in sound, vibration, oil cleanliness, or any other abnormal operation is acceptable cause to 

shut down the operation of the load cell and motor (lubricant pumps should remain on until all motion is 

stopped). A thorough inspection of all components should be conducted before continuing the test. All 

variations should be noted in the test notebook. If there is a shutdown due to mechanical instrumentation 

or lubrication problems, then follow step 9. A preliminary list of computer-monitored conditions follows, 

however there can be additions or subtractions as needed. Each of these conditions has warning levels and 

shutdown levels enabling the test to run unattended. 

• Loss of oil pressure (pumps failed)

• Oil differential pressure is too high across filters (filters clogged)

• Chip detector goes low (ferrous material is in the oil)

• Bearing temperatures are greater than 190 deg F 

• Oil flow is too low 
108

• Oil pressure is too high at the manifold (manifold orifices clogged)

• Vibration is over limits in any of three planes 

• Particle counters indicate high particulate count 

• Torque loss occurs 

• Power loss to test container

• Bulk oil temperature is too great 

• RPM of drive system is out of test limits

• Visual information from the webcam that indicates shutdown is needed

Start-Up Procedure

This discussion might duplicate some steps in the procedure described above, but it is followed when 

restarting the test equipment.

1. Ensure that the oil temperature is 150°F. 

2. Activate the oil circulation pumps. Circulate the lubricant until the oil cleanliness reaches -/15/12. 

With the gearbox coverplate removed, oil should emerge from every bearing and the gear mesh 

nozzles. A lack of oil flow indicates a clogged orifice (lubricant pressure also can be elevated).

3. Energize the motor (motor operation is not possible if the oil flow rate is too low.)

4. Once the motor is operating at the desired speed, gradually apply the load to the system. 

5. After loading has reached the test parameter, confirm that operation of all components is smooth and 

consistent. Log the test time. 

Shutdown Procedure

Smoothly remove the load from the system. In the case of a catastrophic failure, immediate shutdown 

might be required. 

1. Once the load is removed, log the test time. 

2. Smoothly reduce motor speed until the motor stops. 

3. Stop main lubrication pump (which feeds the distribution manifold). During a normal inspection, the 

circulating pumps can remain in operation at the user’s discretion.

4. For extended periods of inactivity, shut down the lubricant heaters and circulating pumps. 

3-Hole Test—Detail of Test

The discussion of the Micon-108 test noted that a four-square type test required the test gearboxes (or 

components) to have the same ratio. This condition created limitations for either the Convoloid or the 

involute gear sets, therefore the “best” involute could not be tested against the “best” Convoloid. Further, 

as center distance increases for gearing, the torque requirements increase dramatically. The chosen center 

distance size for the 3-Hole Test was 4.0 in to allow reasonable testing size without requiring too much 

torque. Related components (couplings, torque actuators, torque transducers), however, are difficult to fit 

into a small space. With this in mind, the 3-Hole Test gearbox and test were designed (Figure A.17).
109

Figure A.17. Basic layout of the 3-Hole Test

Having two gear sets within the test gearbox means that the distance between the shaft extensions is twice 

the operational center distance, thus allowing much more room for other components. Also, the final ratio 

between the shaft extensions for both Convoloid and involute gearboxes is 1:1. As long as the gearing is 

designed to fit into the gear housing, it can be optimized for the test loads. For this test the individual gear 

ratios were approximately 2.2:1. The configuration of the test fixture is discussed in detail above. Some 

testing of the gearing, however, occurred during this time.

After assembling the Convoloid gearing into gearbox 2, a minimal torque was applied as the motor was 

energized to a low speed. Subsequent inspection of the gear teeth revealed acceptable light-load wear 

patterns on the tooth flanks (the flanks had been painted with machinist dye). Figure A.18 and Figure 

A.19 show representative no-load contact on the involute and Convoloid gearing. After the lubrication 

and shaft failure in the Convoloid gearbox, involute gearing was assembled into both gearboxes. All 

subsequent testing occurred using this configuration. 
110

Figure A.18. Representative involute no-load contact

Figure A.19. Convoloid no-load contact

The original test load of 6,300 lb-in yielded a calculated life (L1) for the involute gear set of 42 hours. 

Although this time frame (less than 2 days of 24-hour operation) sounds sufficient, it only represents the 

fact that out of 100 gear sets only 1 possibly would fail in that time. When looking at the average life 

(L50), the expected life is in excess of 24,000 hours (nearly 3 years of test time). Assuming that the 

gearing and materials are of good quality, the average test cycle is prohibitive. Also, at this load the 

projected average life for the Convoloid gearing is more than 700,000 hours. 

Analysis of potential test loads that would shorten the anticipated time-to-failure for surface fatigue yet 

keep the chance of a bending failure remote initially resulted in a test load of 7,000 lb-in. At this load, the 

L50 life for the involute gear set was 9,400 hours—more than 1 year. The L1 life was 16.4 hours. Further 

refining of the analysis resulted in a test load of 7,300 lb-in which gave an L50 life of 6,570 hours (L1 = 

11 hours). The resulting involute bending life is 45,000 hours. For comparison, at the same load 
111

conditions, the Convoloid gearing has an L50 contact fatigue life of 190,000 hours and a bending life of 

more than 300 E6 hours. 

It is important to remember that the bending life must be much greater than the surface durability life due 

to the operation of the test fixture. Once a surface failure is observed and documented, the loading is

reversed (with the same gear set) thus causing contact on the previously unloaded flank. The test is

operated until a surface failure occurs. The test load was chosen assuming the bending L10 life had to be 

greater than twice the surface durability life L50 life. Also, bearing lives and shaft strength considerations 

must be kept in mind. At the 7,300 lb-in load, the involute gear set bending life probability (at twice the 

surface durability life) is ~ 6%. The limiting bearing is the intermediate bearing (recently reduced in size 

to accommodate a larger shaft diameter); however, the shortened catalog life should be offset by the ready 

supply of cool-clean lubricant. If necessary the bearings can be replaced, allowing the test to continue.

Upon installation, the no-load contact pattern for the involute gearing proved to be ideal (Figure A.20). 

Also shown in Figure A.20 is the lubricant supply nozzle for one side of the gear mesh. A similar nozzle 

is positioned below the mesh, thereby assuring oil application at the inlet and outlet of all gear meshes. 

Proper oil pressure (25 psi) is necessary for these nozzles. By the end of the month, the 3-Hole Test 

fixture was nearly operational with good full-load (7,000 lb-in) contact (Figure A.20). Operational speed 

is 2,200 rpm.

Figure A.20. No-load contact pattern—involute gear; lube nozzle
112

Figure A.21. Example of the full load (7,000 lb-in) contact—involute gears

Initial testing for shakedown purposes used a load of 7,000 in-lb which is less than the final approved 

amount of 7,300 in-lb. Early in the test time, however, micropitting was observed in all the loaded flanks 

with more severe wear appearing on the reducer gear sets as compared with the increaser gear sets (see

Figure A.23, Figure A.24). This wear was observed under the existing load, therefore it was decided that 

the test load would remain at the lower level. The operating oil temperature was 155°F—160°F. 

Figure A.22. Involute pinion 1—reducer operation (pinion driving gear)—showing extensive micropitting on 

both the addendum and dedendum
113

Figure A.23. Involute pinion 2—increaser operation (gear drives pinion)—shows a smaller amount of 

micropitting on the pinion (arrow) as well as on the mating gear

Appearance of the micropitting in the involute gear sets caused some concern. Analysis of the load 

conditions, lubricant, and lubricant temperature determined that the test was operating in lubrication

regime between I and II. At this condition there is expected to be some metal-to-metal contact because the 

lubricant film thickness is not sufficient to fully separate the gear flank surfaces. The temperature for the 

test was specified to be 160°F. Although the film thickness was judged to be much too thin for proper 

operation, it was thought that some knowledge could be gained from this test. Subsequent tests will 

reduce the temperature to 140°F, raising the lubrication regime to the region between II and III—thereby 

nearly having a film thickness great enough to fully separate the metal surfaces.

Early in testing, a failure was reported in all of the test gear sets. The test time was 162 hours at 

7,000 lb-in torque, 2,200 rpm, and with an oil temperature of 155–160°F. This comprises approximately 

2.1 times 107 cycles. Although this time is greater than the calculated L1 life of 16 hours, it is nowhere 

near the average life of 9,440 hours. The sudden appearance of surface failures in all gearing in both 

gearboxes was questioned immediately. Initially, the cause was thought to be an abrupt fatigue failure, but 

the appearance of the damaged tooth flanks is not consistent with a classic fatigue case (see Figure A.24). 
114

Figure A.24. Involute (1) reducer pinion surface wear (Test 2)

Instead of classic macropits where material is removed due to subsurface microcracks caused by fatigue, 

there is distinct surface distress that appears to be scuffing (where the contacting tooth flanks weld 

together, then tear away). Investigation into the circumstances surrounding the failure revealed that during 

operation, vibrations during installation of the air conditioner support brackets caused one of the fuses in 

the main circuit breaker to come loose. This initiated an abrupt shutdown of the entire system in which the 

motor stopped while the system was under full load. All other equipment also lost power (e.g., computers, 

lube pumps). 

After the fuse was replaced (and blown fuses in the motor breaker were replaced), the test was restarted. 

After approximately 10 hours of operation, the system was shut down due to high vibration and excessive 

noise. Subsequent inspection of the gearing revealed damage to all gear flanks. It is thought that some 

damage to the gear teeth occurred during the abrupt shutdown and this caused premature wear. Once the 

surfaces were distressed (and/or distorted), additional wear occurred rapidly.

The wear occurred very rapidly as inspections less than 10 hours prior to the loss of power revealed no 

metal particles on the chip detectors and a reduction in the advance of micropitting on the tooth surfaces. 

A normal fatigue situation reveals gradual increases in macropits, beginning on the pinion surfaces 

because these elements see more cycles than the mating gears. This information added to the hypothesis 

that some tooth wear occurred during the shutdown. After the unexpected shutdown and before the final 

hours of operation there were no additional inspections of the gear teeth. 

Examination of test data was not possible because the test computer was not yet configured to gather data 

or to set off alarms for out-of-range conditions. Manual checks of the chip and vibration sensors, 

however, revealed that a shutdown condition existed—which would have been noticed if the computer 

monitoring had been operational (see Figure A.25). Evidently, modifications to the program to allow for 

emergency shutdown and notifications removed the previous settings, causing test personnel to operate in 

a near-manual mode. All data was manually gathered and recorded.
115

Figure A.25. Gearbox 1 chip detector after gearing failure; note material crossing the gap

In addition to updating the computer control for failsafe operation and notifications, controls were added 

for torque adjustment. It originally was thought that the torque setting would be fixed for the duration of 

the test. Temperature changes and other unknown variables, however, have required the hydraulic 

pressure for the torque actuator to be adjusted periodically. Without changes the torque values appear to 

vary by nearly 10%. Before the automated system was operational, the personnel who monitor the test 

performed this touchy function—which requires a “dwell” time after each modification.

Operation of the test fixture is necessary for the computer configuration, therefore it was decided that 

testing would continue with load on the opposite gear tooth flanks. With the previous test, the depth of the 

damage on the gear tooth flanks was deemed to not be detrimental to the bending strength of the gear 

teeth. The calculated L1 bending life at 7,000 lb-in is greater than 150,000 hours. Prior to operation of the 

fixture, the oil was completely drained and filtered twice through a 3-micron filter and the reservoir was 

wiped clean prior to being refilled. The chip detectors also were cleaned to prevent a false positive. 

Finally, the fixture lube system was cycled at operating temperature for more than 1 hour. 

As a modification of the test procedure, any time a “non-normal” situation occurs with the test equipment, 

all gearing as well as the chip detectors are to be inspected for wear and particles. A subsequent 

inspection also should be conducted after 5 to 10 hours, to ensure that no unexpected wear begins on the 

gear teeth. If there is damage then the test on those tooth flanks could be considered void. Additionally, 

because this test is to determine the surface fatigue life of the involute and Convoloid gearing, an 

appropriate lubricant film thickness is necessary. The original oil temperature of 160°F was chosen to 

approximately duplicate the operating conditions inside a wind turbine gearbox. This temperature, 

however, does not allow the lubricant (ISO 320 viscosity) to develop enough of a film thickness to 

separate the metal surfaces of the gear teeth (an oil regime between I and II). Therefore, subsequent 

testing occurred with an oil temperature of 140°F that should allow operation near oil Regime III. 

Operation of the test continued with 100% supervision and manual data gathering with a test load of 

7,000 lb-in torque and 2,200 rpm motor speed. 
116

After 86 hours of operation (Test 3) a significant amount of vibration was noted, prompting shutdown of 

the test fixture. Upon inspection, pinion 1 in gearbox 2 was found to have significant macropitting (see

Figure A.26 through Figure A.28). The origin of the pitting appears to fall at the pitch line of the tooth 

flank, which is consistent with the area of highest calculated stresses for involute gearing. This pinion was 

operated as a reducer where the pinion drives the gear. Note that there is frosting (micropitting) in the 

dedendum (near the tooth root) of the tooth flank. Although, in some cases this phenomenon has been 

thought to precipitate a macropitting failure, at the pitch line where the macropitting originated no 

micropitting is present. Therefore, the macropitting can be considered a classic fatigue failure of the type 

expected in this test. It is important to note that only 10 hours of operation occurred between inspections. 

There was no hint of damage prior to the high vibration level (manually monitored at that time) that 

prompted cessation of the test.

Figure A.26. Massive macropitting on pinion tooth

Figure A.27. Additional pitting flank (Test 3, gearbox 2, pinion 1)
117

Figure A.28. Further pitting and opposite flank scuffing damage from Test 2

The definition of macropitting failure for this test has been set as 1.5% of the operational tooth flank. For 

the involute gearing this translates into a circular pit with a diameter of approximately 0.10 in. The 

Convoloid gearing has a larger contact area, thus the approximate circular pit diameter is ~0.12 in. In this 

failure, the amount of pitting greatly exceeded the definition. It is expected that macropitting will progress 

from small pits to larger ones, and that with the increase in pitted area the measured vibrations will 

increase. This increase in vibration levels will be used for automatic shutdown of the test rig, however, at 

the time of Test 2, the test parameters and warnings had not yet been set up. 

Upon close inspection of the mating gear, no detrimental wear was observed. Therefore, once the failed 

pinion was replaced, testing continued (Test 4) with loading on the same flanks. As noted, the number of 

load cycles for all parts is logged so that, upon failure of a given part, the total number of load cycles can 

be charted. Comparisons between the data scatter for the involute gears versus the Convoloid gears should 

show a distinct increase in component life for the Convoloid gearing. 

Control of the test environment is necessary to reduce as many outside variables as possible, and this was 

finally accomplished with the implementation of the computer-controlled monitoring and control system. 

Up to this point, data values had to be inspected individually using the computer, with the values being 

manually written and no warning or shutdown system in operation. Now the software has been completed 

to allow full monitoring and data collection as well as torque control and alarms (lights and e-mail 

notifications).

In the past, torque control has been a manual operation with the operator watching variation in the torque 

readings and adjusting the torque potentiometer accordingly. The computer now controls the torque and 

holds values much closer to desired values (less than ±.01%) than was possible before. The torque being 

applied to the test remains at 7,000 lb-in. The test speed is 2,200 rpm. 

Test 4 was terminated after 18 hours, due to another surface fatigue failure. Figure A.29 and Figure A.30 

show the failure—again, a pinion operated as a speed reducer, this time in gearbox 1. Here the surface 
118

damage was found during a routine inspection. The amount of damage is indicative of the definition of 

failure. The total test time for this pinion was 104 hours, or 13.7 million cycles. 

It is important to note the extensive micropitting on the tooth flanks. For these parts, it occurs almost 

exclusively in the dedendum of the gear tooth (below the pitch diameter). There has been much discussion 

in the gear industry concerning whether the type of action (speed increaser versus speed reducer) affects 

the appearance of micropitting, or whether it is entirely a result of load, lubrication, and surface finish. 

During testing, it appears that there is a difference due to the type of action. Both of the failed pinions are 

loaded as speed reducers (the pinion drives the gear), and there was extensive micropitting in the 

dedendum of the tooth flanks. The other pinion in the same gearbox with the same number of load cycles 

as the failed pinion in Test 4, yet loaded as a speed increaser (Figure A.31) shows much less wear on the 

active flank. 

Within the gear industry there is much discussion on the effects of gear geometry, how it is loaded 

(increaser versus reducer), and wear potential. A preliminary analysis of the probability for wear on the 

involute gear set (first with the pinion driving the gear, and second with the gear driving the pinion) 

revealed identical specific film thicknesses and similar low probabilities for wear. Outside of the edges 

(tip/root) of tooth contact, the calculated film thicknesses are the same for both analyses. The film 

thickness is a function of the relative curvatures between contacting surfaces, therefore the direction of 

rotation does not appear to have an affect. Additional study into this area is required before definitive 

conclusions can be drawn. 

Figure A.29. Failed involute pinion from Test 4

Figure A.30. Pinion loaded as a speed reducer
119

Figure A.31. The other pinion in the same gearbox loaded as a speed increaser

showing minimal wear

It is thought that micropitting can precipitate macropitting, therefore the presence of this wear is 

important. Contact loads on the surface of a gear tooth flank are directly related to the relative curvature 

between that flank and the mating flank. Standard involute gears have a convex shape; therefore the 

contact patch between the parts is relatively narrow. Further, the convex-convex contact of the gear teeth 

acts to “pump out” lubricating oil, reducing the effective film thickness. In contrast, mating Convoloid 

gear teeth have a convex-concave contact that greatly reduces the relative curvature, creating a wider 

contact patch under load which maintains a larger amount of oil between the surfaces. This increase in oilfilm thickness is expected to reduce or eliminate micropitting in Convoloid gearing. Further, the 

Convoloid tooth form has no contact in the high-stress area of the pitch diameter. The design method 

allows Convoloid teeth to have a consistent contact stress across the entire working flank, which is in 

contrast to the varying stresses in involute gear teeth.

Another aspect that has generated interest is determining if the life of a gear is affected by whether the 

gear component is the driving element (pinion for a speed reducer and gear for a speed increaser) or the 

driven element (gear for a speed reducer and pinion for a speed increaser). To this point, all failures have 

been of the driver element (pinions loaded as an increaser), and early, small macropitting is visible on a 

driving gear loaded as an increaser.

An unexpected failure was encountered (Test 5) involving a tooth breakage in one of the gear teeth 

(Figure A.32). This component originally had been loaded on the opposite flank prior to the scuffing 

failure of Test 2, therefore it was thought that the tooth had failed in bending. Further inspection revealed 

that the hardened outer case had separated from the softer core, causing the tooth to break. This failure—

although not unknown—is very rare and is not included in gearing-life calculation procedures. The gear 

had operated for 163 hours in the current load direction and 174 hours in the opposite direction (a total of 

20.3 million cycles). This time frame represents only a 2.2% probability of failure in bending. A detailed 

inspection of the hardness spectrum and metallurgy was conducted by Northwest Laboratories, Inc. 

(Seattle, WA) to determine if there was material or heat treatment weaknesses in the gear. One theory was 

that the case depth for this part could be less than desirable for the load being applied. 

According to ANSI/AGMA2001-D04, there are three methods that can be used to determine the amount 

of effective case depth for carburized gearing. Two are specified by means equations, in Figure 13 in 

Section 16 of ANSI/AGMA2001-D04, that are related to the tooth size (diametral pitch) only and equate 
120

to values for the effective case depth, he min , of 0.0225 in and 0.0297in for normal and heavy case values 

respectively. It is recommended, however, that the operating effective case depth should be determined by 

using the actual contact stresses expected (equation 43 of the standard), resulting in a value of 0.0262 in. 

Inspection of the broken part showed that the actual case depth was 0.0339 in—well within effective case 

depth specifications. Also, no metallurgical abnormalities were observed during a microscopic inspection 

of sections of the gear tooth. Therefore this broken gear remains a “random event”–type failure and is not 

included in the failure summaries.

Figures A.32. Broken involute gear tooth from Test 5; failure was due to case/core separation

Testing continued after replacement of the damaged gear with the last part in inventory. Some 

adjustments of the computer-control system were necessary to assure a smooth application of torque and 

to help isolate the three axes of vibration. Upon startup, the computer control smoothly ramps up to full 

test load over a period of 30 minutes. This assures full lubrication and no torque spikes that inadvertently 

could damage the gearing. Also, specific set-points were programmed for the accelerometers in each axis 

of the vibration. During previous testing (the Micon 108 test), it was found that vibrations in the minor 

planes showed early wear much sooner and more accurately than the high-vibration planes. The vibration 

in these planes varies after each part change (due to the small magnitude), therefore a different alarm 

point must be set for each test. After the nominal operating vibrations are determined, the alarm point for 

each vibration axis is set (expected to be 50% higher than the nominal). In this manner, because there is 

no all-encompassing set value, the test can be more sensitive to early gear wear. It is hoped that a 

correlation between the system vibrations and early gear flank wear amounts can be developed. 

After a short time of testing, the third macropitting failure in the involute gearing occurred. Once again, 

the failed member was a driving pinion—this time with 82 hours of operation. This part had replaced an 

earlier failed pinion. To date, the only macropitting failures have been with driving pinions with 86, 104, 

and 82 operation hours. The other involute pinions (being driven by the gear) have survived for more than 
121

168 hours with minimal wear visible. A small macropit is visible on a driver gear, however this failure 

has not reached the test definition of failure. All remaining gear flanks show minimal wear. 

A summary of the relevant tests is shown in Figure A.33. The graph excludes the broken gear (Test 5) and 

early setup cycles (Test 1) where there was no failure in the involute gearing. A previous failure (Test 2, 

which was due to a loss of lubricant), is included to show that these parts had exceeded the current failure 

lives. Therefore, it is thought that subsequent tests should indicate longer lives than what these three tests 

reveal.

Figure A.33. Summary of involute gear failures in the 3-Hole Test

Testing currently has been suspended due to lack of funding. There are eight sets of involute gearing and 

one set of Convoloid gearing available for testing. An additional 6 sets of Convoloid gearing await 

confirmation of the micropitting protocol before having a final tooth grind. The anticipated time frame for 

testing, based upon gearing components currently in process, is expected to depend primarily on the 

number involute pinions available because they have the shortest expected lifespan. Assuming both tooth 

flanks of each component are able to be loaded in turn and a component lifetime of approximately twice 

the current failure rate, the existing parts being manufactured allow for approximately nine months of 

operation. During this time, approximately 24 data points should be gained for the involute gearing and 

approximately 10 data points should be gained for the Convoloid gearing. If testing follows fatigue 

theory, then the time frame will be greatly lengthened. 0.00E+00 5.00E+06

1.00E+071.50E+072.00E+072.50E+07

Driver Pinion (Test 2)

Driven Pinion (Test 2)

Driven Pinion (Test 2)

Driver Pinion (Test 2)

Driver Pinion (Test 3)

Driver Pinion (Test 4)

Driver Pinion (Test 6)

FULL LOAD CYCLES

LOADED FLANK IDENTIFICATION.

Involute 3-HOLE TEST Results - 2008-Oct-3
122

Unless additional involute pinions and gears are made later in testing, it is expected that some of the 

Convoloid parts will not be used by the time the last involute part has failed. To this point, all testing has 

coincided with configuration of the test fixture and only has involved involute gearing after the original 

Convoloid gearing was damaged due to a lubricant pump failure. Data gathered from the involute gear 

failures is expected to act as the control for the Convoloid testing. These data points are expected to 

confirm current involute gear theory.

Implementation of the full test—involute versus Convoloid—is expected after the Convoloid gearing all 

have been re-ground to achieve an appropriate no-load contact. Only one set of Convoloid gearing 

previously has been ground to act as a qualification of the grinding method and to ensure that premature 

micropitting wear does not appear on the Convoloid profile (as was experienced in the Micon 108 test). 

Estimated Test Time—Continuation of 3-Hole Test

3-Hole Test

Estimated Time of Test with Existing Components

as of 9/8/2008 

As preparation for full testing of the 3-Hole Test project, it is necessary to estimate how long testing can 

continue with the parts currently in inventory or in the manufacturing process. At the time of this writing, 

the parts available for this test (both finished and in process) are as follows.

• Convoloid Pinion 1—1 complete, 1 possible scrap at Bradford Inst., 1 in process at Carnes-Miller, 

6 in process at The Gear Works (TGW)

• Convoloid Pinion 2—1 complete, 1 possible scrap at Bradford Inst., 1 in process at Carnes-Miller, 

6 in process at TGW 

• Convoloid Gear—2 complete, 4 in process at Carnes-Miller, 7 in process at TGW

• Involute Pinion 1—8 in process at TGW

• Involute Pinion 2—8 in process at TGW

• Involute Gear—17 in process at TGW 

The anticipated test time is dependent upon the length of time necessary for components to fail. It is 

expected that components will fail with a surface fatigue (macropitting) mode, however experience has 

shown that unanticipated failures are possible. Further, a minimum amount of time is required for 

changing-out failed components and reassembling the fixture with new parts. The estimation of life for 

test components is based upon both current testing experience as well as fatigue theory. There have been 

failures earlier than theory predicts, therefore the estimated life must reflect that situation. Due to theory, 

however, it is assumed that the average life of test components will be longer than in previous tests. If 

components last much longer—to their calculated average life, for example—then testing will require 

additional time.

Due to the reduced number of cycles, the gears have a longer life (in hours) than the pinions. Therefore 

the pinion life becomes the limiting factor. Also, because there is a much greater bending life, both flanks 
123

of the pinions are able to be loaded with minimal risk of having broken teeth. Lastly, as one component 

suffers a surface failure, early testing has revealed that no adverse damage occurs to the mating flank. 

This is not assumed, however, and a thorough inspection is conducted after every failure. If no adverse 

damage occurs to the mating flank, then only the failed component must be replaced—allowing for 

additional test points.

As components fail they must be replaced. Provisions within the test fixture allow the torque to be 

reversed without modifying the gearboxes or the fixture. Therefore, in many cases, two tests can be 

conducted before it is necessary to shut down the fixture and replace parts. From previous testing, it is 

estimated that the changeover requires 37.5 man-hours. Replacing damaged components requires a 

number of steps. 

1. Removal of the gearbox from the test fixture

2. Disassembly of the affected portion of the gearbox

3. Cleaning of components 

4. Reassembly of the gearbox with replacement of the damaged component(s)

5. Installation of the gearbox onto the test fixture

6. Balancing couplings to reduce vibrations 

7. Heat test of fixture oil and thoroughly circulate it to ensure clean lubricant

8. Start the new test

Following is a summary of the anticipated length of the test with current components, based upon an 

anticipated failure rate. Variations in these failure rates are expected.

Involute

The average life of all failed components (including those that lasted until the lubrication incident) was 

148 hours, with the shortest being 86 hours. It is thought that the times of these failures are on the low end 

of the spectrum because all the parts of Test 2 lasted until a power outage caused loss of lubricant and 

resulted in a scuffing failure (164 hours) before loading the opposite flank. Therefore, a value of 220 hr is

used as the anticipated life. According to the life calculations, this represents an L10.5 life (10.5% 

probability of failure). With the average surface contact life (L50) at 6,460 hours, it would be expected that 

testing should progress well beyond the L10.5 level. That has not been the case in early testing, however. 

The corresponding bending life (with double the L50 contact cycles—because both flanks will be loaded) 

results in an L4.6 life (4.6% probability of failure). Based upon the life of the pinions, the approximate test 

time becomes the following. 

Test Time = 220 hours * 8 pinions * 2 flanks = 3,520 hours 

The amount of time required for changing-out failed components is reduced somewhat due to the ability 

to reverse loading in the test. Unexpected situations could arise, however, causing some part changes.

Based on the available components, the approximate test time required for replacing failed components 

(assuming an active test time of 3,520 hours) becomes the following. 

16 total pinions, 1 pinion failure at the end of the test + 8 gear changes + 5 unexpected changes = 

30 gearbox changeovers
124

28 changes * 37.5 hours/changeover = 1,050 hours 

Convoloid

Using the same failure definition as for the involute gearing, the life of the Convoloid pinions is 320 

hours. The average surface contact life (L50) is calculated to be 190,000 hours. The corresponding bending 

life (with twice the L50 contact cycles) results in an L7.8 (7.8% probability of failure). Initially, this seems 

a higher probability than for the involute, however the Convoloid gears see a number of cycles that is 

29 times greater. 

With the life of the Convoloid pinions expected to exceed that of the involute pinions and with the usable 

number of parts being equal, the only addition to the total test time will be the time it takes to change over 

the Convoloid gearbox. Based upon the life of the pinions, the approximate number of pinions to be used 

during the involute test time becomes following. 

Number of Convoloid pinions = 3,520 hours / 320 hours / 2 flanks = 5.5 ~ 6 pinions 

Therefore, potentially 6 of each pinion will be used during the time it takes to fail all 8 of the involute 

pinions. In that time it is expected 3 gears for each pinion will fail (due to the reduced number of cycles 

on the gears). This leads to the number of changeovers and thus the amount of time necessary for 

Convoloid changeovers. 

(12 total pinions + 6 gear changes + 4 unexpected changes) = 22 gearbox changeovers 

22 changes * 37.5 hours/changeover = 825 hours 

Therefore, the total anticipated test time is 3,520 + 1,050 + 825 hours = 5,395 hours. Allowing 20% for 

contingencies, this leads to a total consumed time of 6,474 hours (approximately 9 months). 

Early Test Results

3-Hole Early Test Results

2008 Aug 26 

• Test Components: 16 x 35 involute gearing, 7 NDP, 20° NPA, 18.697° HA, 1.450 in FW, 

4.00 in CD

• Test Conditions: 

• Test Load: 7,000 lb-in pinion torque 

• Contact Stresses: 218,715 psi calculated per ISO6336 (AGMA Contact Stress Number 

232,994 psi) 

• Calculated L1 life: 4.29E7 cycles (325 hours) per ISO6336; 5.28E6 cycles (40 hours) per AGMA: 

2001 
125

Figure A.34. Involute 3-hole test results

Failure Modes

• Test 2: Severe scuffing due to loss of lubricant 

• Test 3: Extensive macropitting on numerous increaser pinion flanks 

• Test 4: Macropitting on two increaser pinion flanks 

• Test 5: Test ongoing—no damage on pinions, slight pitting on one reducer gear 

• All components showing macropitting are the driving member 0.00E+00 5.00E+06 1.00E+07 1.50E+07

2.00E+072.50E+07

Reducer Pinion (Test 2)

Increaser Pinion (Test 2)

Increaser Pinion (Test 2)

Reducer Pinion (Test 2)

Increaser Pinion (Test 3)

Increaser Pinion (Test 4)

Reducer Gear (Test 5)

(ongoing)

Reducer Pinion (Test 5)

(ongoing)

FULL LOAD CYCLES

LOADED FLANK IDENTIFICATION

Involute 3-HOLE TEST Results
126

3-Hole Test Life Analysis

3-Hole Test: Involute stresses at 7,000 lb-in

Figure A.35. Contact probability (L9) that matches current failure lives; bending L1 life is much higher

(for 2X contact cycles)

Figure A.36. Calculated L50 contact life and the corresponding bending life (L4.6) 
127

Low Wind Speed Turbine Baseline Turbine— 

Baseline Operating and COE Parameters

Rev 1. 11/09/2004

The following describes the operating conditions and parameters used to establish the 2002 LWST 

Baseline Turbine cost numbers. These operating conditions and numbers will be used for the validation of 

TIO projections and for improvement forecasts for LWST subcontractor’s whenever appropriate. Most of 

the detailed component numbers are based upon the WindPACT study work performed by GEC. There 

have been minor adjustments to match these numbers to those selected by program management for the 

baseline conditions. Where those conditions vary, they have been noted below. Accompanying this 

document will be two spreadsheets. One spreadsheet is an AEP calculator that was created by Lee 

Fingersh (National Wind Technology Center). This calculator allows adjustments for wind speed, weibull 

shape factor, shear, rotor diameter, hub height, air density, rotor Cp, and various efficiencies. This 

spreadsheet should be used when analyzing potential improvements in performance and operating 

conditions. The second spreadsheet is a COE calculator. It uses the baseline cost information developed 

by GEC under WindPACT with changes as noted below. This spreadsheet should be used when analyzing 

the impact of changes in component costs on overall COE. It also has cost elements for O&M, Levelized 

Replacement Cost and Land Lease Cost.

Operating Conditions and Parameters

WindPACT (1999) Baseline Turbine (2002)

Rotor Diameter 70 m 70 m

Rating 1,500 kW 1,500 kW

Hub Height 65 m 65 m

Operating Wind Class 4 5.8 ms at 10 m 5.8 ms at 10 m

Weibull K Factor 2 2

Base Wind Shear 1/7 (0.143) 1/7 (0.143)

Altitude 0 m 0 m

Air Density 1.225 kg/m^3 1.225 kg/m^3

Rotor Cp (This Cp has been reduced from the WindPACT 

study to more closely match the projected Cp of a machine 

of this size in 2002, as based on survey data.)

0.5 0.47

Conversion Efficiency (This conversion efficiency is 

actually represented as an efficiency surface in the spread 

sheet, and matches the profile of the WindPACT studies.)

0.95 0.95

Soiling Losses (Soiling Losses have been increased slightly 

to match with the combined losses used for projecting the 

2002 Baseline)

2% 3.5%

Array Losses (The product of the conversion efficiency, 

soiling losses and array losses is a reduction in AEP of 13%. 

This matches the losses for the 2002 Baseline before 

application of the availability.)

5% 5%

Availability (Availability has been increased from the 

WindPACT 95%, to the 98% used for the 2002 Baseline 

Turbine projection. This more closely matches reported 

project numbers for recent installations.)

95% 98%
128

Cost of Energy Adjustments

The following adjustments have been made to the COE spreadsheet to match the numbers with those 

derived for the LWST 2002 Baseline Turbine. 

Manufacturing Uncertainty 

WindPACT (1999) $0 

Baseline Turbine (2002) $162,000 

A factor called manufacturing uncertainty has been added to the initial capital cost for the turbine; this 

number has been set at $162,000. It is included as an added markup to make WindPACT capital cost 

numbers consistent with a wide range of reported costs per kilowatt for large (100 MW and larger) 

projects reported in the 2002 timeframe. The WindPACT component cost data was developed based on 

quotes from vendors and cross checks with other industry data, where available, on a component-bycomponent basis. It is believed, however, that due to less than optimum production conditions, the advent 

of newer equipment, starts and stops in production due to uncertainties in the Production Tax Credit, 

exchange rate risks, and less than ideal timing of project starts that manufacturer costs or mark-ups are 

running above those assumed in WindPACT studies. As the LWST project proceeds this number could be 

reduced as better cost estimates are obtained. 

Tower Costs

WindPACT (1999) $183,828 

Baseline Turbine (2002) $101,000 

The initial WindPACT tower was based on an 84-m hub height. For the Baseline Turbine, this hub height 

has been reduced to 65 m, consistent with the majority of recent projects. This has reduced the baseline 

tower costs to $101,000 from the original estimate of $183,828. 

O&M

WindPACT (1999) $0.008/kWh

Baseline Turbine (2002) $0.007/kWh 

The WindPACT O&M cost number was fixed at $0.008 per kWh. This was intended to limit O&M being 

varied during WindPACT studies, because these studies primarily focused on determining the impact of 

component design changes. For the Baseline Turbine the O&M number has been reduced to $0.007 

per kWh (based on an estimate of $30,000 per turbine), to more closely match recent reports. For 

simplicity, the COE spreadsheet has reduced this calculation to $20 per kW per year. It is important to 

note that O&M is a tax-deductible expense. In the final COE calculation in the spreadsheet, the O&M 

number is multiplied by 0.6 to take into account the tax-deductible nature of the expense. 

Levelized Replacement Cost

WindPACT (1999) $15/kW/turbine 

Baseline Turbine (2002) $10.70/kW/turbine

Long-term replacement and overhaul costs from WindPACT were set at $15 per kW per turbine. For the 

2002 Baseline Turbine this number was lowered to $10.70 per kW. 
129

Land Cost

WindPACT (1999) $0 

Baseline Turbine (2002) $0.00108/kWh/turbine 

For the WindPACT study it was decided to include the lease cost of land in the fixed-charge rate, along 

with several other fixed costs. For the 2002 Baseline, the land-lease cost has been entered as a separate 

item in the spreadsheet and set at $0.00108 per kWh per turbine.

Fixed Charge Rate 

WindPACT (1999) 10.6% 

Baseline Turbine (2002) 11.85% 

The WindPACT fixed charge rate (FCR) was set at 10.6%. This number was established at the beginning 

of the WindPACT project in late 1999. For the Low Wind Speed Technology project this number was 

adjusted to 11.85% to be more in line with data for projects at that time (2002). Additionally, the FCR 

was updated as a result of efforts to more closely align the pro forma cash-flow spreadsheet methodology 

with industry practices. The FCR is imputed from a standard case using the cash-flow spreadsheet. The 

FCR reflects finance charges and cost of money as well as other factors and in reality fluctuates over the 

years. For purposes of comparing competing technologies, however, it is necessary to freeze this number 

as has been done for LWST. 
130

Appendix B. Baseline Cost of Energy Comparison Sheet (2002 $)

Baseline COE Comparison Sheet (2002 $)

Baseline Turbines:

• 108 kW—3-Bladed Upwind/Fixed Pitch Stall Controlled (from Wally T. report)

• 750 kW—3-Bladed Upwind/Variable Pitch Controlled (from WindPACT Turbine Rotor Design 

Study, June 2000 to June 2002) 

• 1.5 MW—3-Bladed Upwind/Variable Pitch Controlled (from WindPACT Turbine Rotor Design 

Study, June 2000 to June 2002) 

• 3.0 MW—3-Bladed Upwind/Variable Pitch Controlled (from WindPACT Turbine Rotor Design 

Study, June 2000 to June 2002) 

• 5.0 MW—3-Bladed Upwind/Variable Pitch Controlled (from WindPACT Turbine Rotor Design 

Study, June 2000 to June 2002) 

Rating (kW) 108 750 1,500 3,000 5,000

Baseline

Component

Costs $1,000

Baseline

Component

Costs $1,000

Baseline

Component

Costs $1,000

Baseline

Component

Costs $1,000

Baseline

Component

Costs $1,000

Component

Rotor 19.0 101.90 247.53 727.93 1484.43

Blades 15.0 64.07 147.79 437.46 905.90

Hub 4.0 21.62 64.19 213.03 429.31

Pitch mechanism & bearings 0.0 16.21 35.55 77.44 149.22

Drive train, nacelle 44.8 255.63 562.77 1282.00 2474.26

Low speed shaft 1.5 8.43 19.86 56.26 120.90

Bearings 1.8 3.79 12.32 41.44 101.83

Gearbox 28.0 64.92 150.88 357.22 697.06

Mechanical brake, HS coupling, etc. 1.5 1.49 2.98 5.97 9.95

Generator 3.5 48.75 97.50 195.00 325.00

Variable speed electronics 0.0 50.25 100.50 201.00 335.00

Yaw drive & bearing 2.3 5.27 12.09 28.21 109.71

Main frame 2.5 21.45 63.99 192.12 433.63

Electrical connections 2.7 30.00 60.00 120.00 200.00

Hydraulic system 0.5 3.38 6.75 13.50 22.50

Nacelle cover 0.5 17.90 35.90 71.28 118.68

Control, safety system 13.0 10.00 10.20 10.49 10.78

Tower 18.0 69.66 183.83 551.42 1176.15

Turbine Capital Cost (TCC) 94.8 437.19 1004.33 2571.84 5145.62

Foundations 4.0 34.92 48.51 76.77 108.09

Transportation 3.5 26.59 51.00 253.41 1312.15

Roads, civil works 2.0 44.90 78.93 136.36 255.33

Assembly & installation 5.0 24.37 50.71 112.71 224.79

Electrical interface/connect 7.0 71.30 126.55 224.20 431.50

Permits, engineering 2.5 15.79 32.70 69.87 126.39
131

Rating (kW) 108 750 1,500 3,000 5,000

Baseline

Component

Costs $1,000

Baseline

Component

Costs $1,000

Baseline

Component

Costs $1,000

Baseline

Component

Costs $1,000

Baseline

Component

Costs $1,000

Balance of Station Cost (BOS) 24.0 217.87 388.41 873.31 2458.24

Project Uncertainty 0 0.00 0.00 0.00 0.00

Initial Capital Cost (ICC) 118.80 655.06 1392.74 3445.15 7603.86

Installed Cost per kW (cost in $) 1,100.00 873.41 928.50 1148.38 1520.77

Turbine Capital per kW sans BOS (cost in $) 877.78 582.92 669.55 857.28 1029.12

Levelized Replacement Costs (LRC) 

($10.7/kW) 12.00 8.03 16.05 32.10 53.50

O&M $20/kW/Yr (O&M) 2.16 15.00 30.00 60.00 100.00

O&M ($/kWh) 0.0040 0.0037 0.0035 0.0033

Land ($/year/turbine)($0.00108/kWh) 0.25 2.43 5.20 11.20 19.58

Net 5.8 m/s Annual Energy Production MWh 

(AEP) 228.83 2219.73 4733.53 10265.32 18142.64

Net 6.67 m/s Annual Energy Production 

MWh (AEP) 305.04 2785.45 5840.23 12456.11 21743.67

Net Annual Energy Production MWh (AEP) 

(per WindPACT Study) N/A 2254.463 4816.715 10371.945 18132.994

Fixed Charge Rate 11.85% 11.85% 11.85% 11.85% 11.85%

COE at 5.8 m/s $/kWh (see note 2) 0.1207 0.0437 0.0432 0.0475 0.0570

COE at 6.67 m/s $/kWh (see note 2) 0.0905 0.0349 0.0350 0.0391 0.0476

COE (per WindPACT study)(unknown wind 

velocity-IEC Class 2) 0.0431 0.0424 0.0470 0.0570

Notes:

1. The WindPACT study specified replacement costs at $15.00/kW/year, a 10.6% fixed charge rate, and O&M costs fixed at 

$0.008/kWh.

2. The 5.8 m/s and 6.67 m/s values for the 750 kW and larger turbines were calculated using wind data from “Alternative 

Design Study Report: WindPACT Advanced Wind Turbine Drive Train Designs Study,” and used the same spreadsheet 

Wally Thompson used for the 108 kW.
132

Appendix C. Detailed Task Outline

This is the revised task outline as determined during the July 27, 2005, kick-off meeting.

Main Objectives of the Program

• Decrease in capital cost for given capacity WTG gearbox

• Reduction in size and weight 

• Potential increase in performance and/or efficiency

Accounting Task Codes in parentheses

Assignments—7/27/05 Timeline Responsibility

Task 1. Initial COE Projection—Deliverable 1 26 July 05

1.1 (101) Review of stress comparisons—Micon 108 gearing

1.1.l Review of rating factors assumed

MEI, JRC, AJW

1.2 (102) Establish O&M costs for Micon 108 gearbox WT, EUI

1.3 (103) Projection of rating factors and stress levels to be used in 

theoretical design of similarly rated Convoloid gear system

1.3.1 Project O&M costs as a function of operating stress level

MEI, JRC, WT

1.4 (104) Optimize design of Convoloid box for 108-KW capacity with 

reduced O&M costs

WT, MEI, JRC

1.5 (105) Analysis of cost factors for Convoloid gearbox

1.5.1 Compare to full-size (existing) involute box

TGW

WT

1.6 (106) Calculate COE reduction based on cost, weight, and O&M 

factors for Convoloid box

1.7 (107) Other

Task 2—Kick-Off Meeting—Meet at TGW

2.1 (201) Review plan details 27/28 July 05 BB

2.2 (202) Potential testing activities MEI, TGW, 

DCE

2.3 (203) Future directions with NREL Project Team BB

2.4 (204) Meeting summary BB

Task 3—Gearbox Comparison Tests and Protocols

3.1 (301) Refine Manufacturing Processes

3.1.1 Reconcile grinding wheel dressing techniques for Convoloid 

gearing

3.1.2 Measure and establish fixes for suspected Micon 108 housing 

distortions under 150%+ rated loading

3.1.3 Develop and execute on-machine tooth profile inspection 

protocol for Convoloid forms

3.1.4 Develop and execute off-machine tooth profile and 

comprehensive gear inspection techniques for Micon 108 

Convoloid gears

3.1.5 Disassemble and inspect all Convoloid gears with new CMM 

software

19 Aug. 05 JRC, TGW, 

AJW

10 July 05 MEI, TGW

9 Sept. 05 TGW, JRC, 

AJW

BB, TGW

TGW
133

Accounting Task Codes in parentheses

Assignments—7/27/05 Timeline Responsibility

3.1.6 Regrind and recheck all Convoloid gears until quality level is 

determined to be satisfactory (i.e., at least equal to the involute 

accuracy)

26 Aug. 05 TGW

3.2 (302) Compare stress levels at 200% rated load to insure bending 

strength life far surpasses that for surface durability

26 July 05 MEI, JRC, 

TGW

3.3 (303) Calculate the life of both boxes under their respective loads in 

a surface failure mode

3.3.1 Define “failure”

MEI, JRC, GEI

MEI, JRC, GEI

3.4 (304) Analyze manufacturing sensitivities of Convoloid gearing and 

their effects on rating factors that should be used (at least initially) for 

time to failure calculations

3.4.1 Specific sliding

3.4.2 Absolute sliding

3.4.3 Film thicknesses

3.4.4 Center distance changes

3.4.5 Profile tolerances

3.4.6 Profile shifts

3.4.7 Spacing tolerances:

 3.4.7.1 Individual

 3.4.7.2 Accumulative

3.4.8 Effects of crown

3.4.9 Effects of lead corrections

3.4.10 Gross misalignment

3.4.11 Bearing capacity

3.4.12 Tangential and radial forces compared to involute designs

3.4.13 Possible constraints and/or enhancements that could impact 

the application of Convoloid gears in the future

 3.4.13.1 Transition zone mismatches

 3.4.13.2 Separating loads

3.4.13.3 Transition zone height overlaps

31 Oct. 05 GEI, MEI, JRC, 

AJW

30 Sept. 05 WT, MEI, JRC

26 July 05 TGW, MEI

23 Aug. 05 MEI, TGW

3.5 (305) Relate testing protocol to expected loads for a Class 4 wind 

site

3.6 (306) Full analysis of housings, bearings, shafts, keys, and other 

components to determine safety factors of these components under 

protocol loads and speeds (to preclude premature failure)

3.7 (307) Description of planned test protocol

3.8 (308) Other

Task 4—Testing of Involute and Convoloid Gear Sets

4.1 (401) Instrumentation and data gathering system 26 July 05 DCE

4.2 (402) Parameter limits for automatic shut down 26 July 05 DCE, MEI

4.3 (403) Failure modes and precursors 26 Aug. 05 MEI, JRC, 

TGW

4.4 (404) Photography 20 Sept. 05 MEI

4.5 (405) Noise measuring 20 Sept. 05 MEI, DCE

4.6 (406) Execute test protocol 20 Sept. 05 DCE, TGW, 

GEI, MEI

4.7 (407) Other 18 May 06 AJW, JRC
134

Accounting Task Codes in parentheses

Assignments—7/27/05 Timeline Responsibility

4.10 (410) 3-Hole Test preliminary gearbox design

4.11 (411) 3-Hole Test project oversight

4.12 (412) 3-Hole Test—test fixture

4.13 (413) 3-Hole Test—test procedure

4.14 (414) 3-Hole Test—instrumentation 

4.15 (415) 3-Hole Test—gearbox design

4.16 (416) 3-Hole Test—gearbox component detailing

4.17 (417) 3-Hole Test—gearbox component manufacture

4.18 (418) 3-Hole Test—gearbox assembly

4.19 (419) 3-Hole Test—test apparatus assembly

4.20 (420) 3-Hole Test—execute test protocol

Task 5—Detailed Analysis of Test Results and Predictions

5.1 (501) Comparison of Convoloid versus involute gear drive systems 03 Nov. 05

DCE, MEI, 

JRC, TGW, 

GEI

5.2 (502) Relationship of Convoloid gear performance to AGMA rating 

formulae and factors 2 Dec. 05 MEI, JRC, TGW, GEI

5.3 (503) Relationship of involute gear performance to AGMA rating 

formulae and factors 2 Dec. 05 MEI, JRC, TGW, GEI WT

5.4 (504) Trade-off summary—minimized COE for baseline turbine 

(projection of COE versus surface durability and fillet stress 

compromises)

15 Jan. 06 WT

5.5 (505) Extrapolation of data across a wide range of WTG gearboxes 

(see SOW for details)

30 Apr. 06 JRC, AJW, WT

5.6 (506) Test results report—Deliverable 3 15 June 06 MEI, GEI

5.7 (507) Other

Task 6—Field Testing

6.1 (601) Planning of Field Test Program

6.1.1 Establish number of units of each model

6.1.2 Location of field test(s) 

6.1.3 Develop periodic inspection plan

6.1.4 Instrumentation design

6.1.5 Convoloid gearbox design

6.1.6 Involute gearbox design

Start

2 Jan. 06

Finish

15 Jan. 06

MEI, DCE, 

TGW, EUI, 

GEI

6.2 (602) Build instrumentation 7 Feb. 06 DCE

6.3 (603) Build Convoloid gearboxes 1 Feb. 06 TGW

6.4 (604) Build involute gearboxes 15 Jan. 06 TGW

6.5 (605) Ship gearbox/instrumentation packages to destination 15 Feb. 06 TGW

6.6 (606) Assemble gearboxes and instrumentation packages on site 1 Mar. 06 DCE, EUI

6.7 (607) Execute periodic inspection plan 15 Mar. 06 EUI

6.8 (608) Test results report—Deliverable 4—monthly start 15 Mar. 06 MEI, EUI, GEI

6.9 (609) Purchase long-lead items 3 Oct. 05 BB
135

Accounting Task Codes in parentheses

Assignments—7/27/05 Timeline Responsibility

Task 7—Detailed Design of Scaled-Up Gearbox Using Convoloid 

Gear Sets

7.1 (701) Selection criteria for candidate gearbox Start

15 June 06

TGW, MEI, 

EUI, GEI

7.2 (702) Prioritized list of objectives in scale-up effort 1 July 06 TGW, MEI, 

EUI, GEI

7.3 (703) Match 7.1 with 7.2 15 July 06 TGW, MEI, 

EUI, GEI

7.4 (704) Obtain actual gearboxes and/or drawings of candidates 15 Oct. 06 TGW, MEI, 

EUI, GEI

7.5 (705) Reverse-engineer and design Convoloid retrofit

7.5.1 Analyze Convoloid gear ratings using data developed in (5.2) 

above

7.5.2 Analyze bearing life, performance improvements; compare 

with existing Involute design

7.5.3 For the “selected” involute design using calculated stress 

levels, geometry, materials, and heat treatment, compare potential 

reductions in weight and cost of a “like-rated” Convoloid box

31 Dec. 06 JRC, AJW

7.6 (706) Other 1 Mar. 07 MEI, JRC, 

TGW, GEI

Task 8—Final Report, Deliverable 5 15 Apr. 07 MEI, GEI, BB, 

AJW
136

Appendix D. Gearing Life Analysis—Surface Failure Mode

4 November 2005 

Definition of Surface Failure

Gear industry standards reference: AGMA 912-A04, ANSI/AGMA 1010-E95, ISO 

10825:1995(E/F)

No definitive specification exists within the listed standards that explicitly states when a gear (or gear 

tooth) has failed in a surface failure mode. Additionally, combinations of different failure modes are 

common. Therefore determining the original cause requires years of experience. Unfortunately, many 

determinations are made after a gross destruction of a tooth flank. A summary of the most common 

surface failure modes is included at the end of this appendix. 

The primary surface failure mode for which this study is concerned is a fatigue-type failure known as 

“macropitting.” Therefore great care is required to prevent other failure modes (i.e., those due to 

insufficient lubrication, improper lubrication, mesh misalignment, manufacturing deviations, or improper 

material). The expected life of a surface in fatigue is directly related to contact stress and number of 

cycles.

As a surface is loaded, portions of the surface flex and eventually break away, leaving a “pit” or shallow 

depression in the surface. Most pits are small, but they can grow together and precipitate other failure 

modes—especially tooth breakage. The size of pitting and percentage of a tooth flank that is covered 

determine the definition of failure.

In some industries where transmission noise and smoothness is paramount, very little pitting is allowed. 

These industries specify failure as a very small percentage of the gearing contact face width. Other 

industries are concerned strictly with the transmission of rotational power, and therefore broaden the 

definition to the point of tooth breakage. For the wind turbine industry, the gearbox is an extremely 

important component of the turbine and, as such, must be very reliable. Failure of the gearbox creates 

extreme financial difficulties therefore Genesis defined failure as “any defect that will progress to the 

economic (operational) failure of the wind turbine gearbox.” 

Summary of Types of Common Surface Failure Modes

Pitting (Macropitting)

A contact fatigue failure occurs when fatigue cracks initiate either at the surface of the gear tooth or at a 

shallow depth below a loaded surface. Pits are formed when these cracks propagate back up to the surface 

and material separates. According to AGMA 925-A03, “Damage beginning on the order of 0.5 to 1.0 mm 

in diameter is considered to be a macropit.” Four classifications have been observed: non-progressive, 

progressive, flake, and spall. Additional information and descriptive photos can be found in ANSI/AGMA 

1010-E95, clause 6.1; and ISO 10825, clause 4. 

Micropitting

Micropitting is a localized contact fatigue failure that exhibits a frosted, matte, or gray-stained 

appearance. The dynamic is the same as macropitting; however the pits are very fine. Micropitting occurs 

most frequently on surface-hardened gear teeth. Additional information and descriptive photos can be 

found in ANSI/AGMA 1010-E95, clause 6.2; and ISO 10825, clause 4.1.3. 
137

Wear

This is a non-fatigue failure that is a degradation of a gear tooth involving removal or displacement of 

material due to mechanical, chemical, or electrical action. Specific classifications include adhesion, 

abrasion, polishing, corrosion, fretting corrosion, cavitation, erosion, electrical discharge, and rippling. 

Additional information and descriptive photos can be found in ANSI/AGMA 1010-E95, clause 3; and 

ISO 10825, clause 1.1, 1.2, 1.4, and 1.5. 

Scuffing

This is a non-fatigue failure that involves severe adhesion between mating gear elements causing transfer 

of metal from one tooth surface to another due to welding and tearing. Degrees of scuffing are classified 

as mild, moderate, and severe. Additional information and descriptive photos can be found in 

ANSI/AGMA 1010-E95, clause 4; and ISO 10825, clause 2. 

Plastic Deformation

This is a permanent deformation that occurs when the stress exceeds the yield strength of the material. 

Specific classifications include: indentation, cold flow, hot flow, rolling, tooth hammer, rippling, ridging, 

burr, root fillet yielding, and tip-to-root interference. Additional information and descriptive photos can be 

found in ANSI/AGMA 1010-E95, clause 5; and ISO 10825, clause 3. 
138

Appendix E. Micon 108 Instrumentation Data Points

Micon 108 Lab Test Instrumentation Parameters

Note:

1. Involute gearbox is #1 gearbox (nearest the motor)

2. Convoloid gearbox is #2 gearbox

Recorded in 

Data File Alarms Signal Type

Convoloid oil flow Y Pulses

Convoloid oil temp out Y

Convoloid oil temp in Y 

Convoloid particles Y

Convoloid oil pressure at manifold Y Loss of Oil Pressure DC Volts

Convoloid, outside high speed/intermediate separating 

force Y SG

Convoloid, outside low speed/intermediate separating 

force Y SG

Convoloid, outside, low speed bearing temperature Y Excessive Temperature TC Isolated

Convoloid, outside, intermediate bearing temperature Y Excessive Temperature TC Isolated

Convoloid, outside, high speed bearing temperature Y Excessive Temperature TC Isolated

Convoloid, inside high speed/intermediate separating 

force Y SG

Convoloid, inside low speed/intermediate separating 

force Y SG

Convoloid, inside, low speed bearing temperature Y Excessive Temperature TC Isolated

Convoloid, inside, intermediate bearing temperature Y Excessive Temperature TC Isolated

Convoloid, inside, high speed bearing temperature Y Excessive Temperature TC Isolated

Convoloid reservoir temperature TC

Convoloid gearbox sump temperature Y TC

Convoloid magnetic chip detector Y Particles Detected DC Volts

Convoloid filter pressure drop Y Excessive Filter Pressure Drop

Convoloid X vibration Y Excessive Vibration AC mV

Convoloid Y vibration Y Excessive Vibration AC mV

Convoloid Z vibration Y Excessive Vibration AC mV

Convoloid total vibration (calculated) Y Excessive Vibration calc

Convoloid heat exchanger oil temp in TC

Convoloid heat exchanger oil temp out TC

Convoloid particle count 4u DC Volts

Convoloid particle count 6u High Count DC Volts

Convoloid particle count 14u High Count DC Volts

High speed shaft RPM (motor RPM) Y Pulses

Circulating torque Y SG

Windup force strain gage Y SG

Windup force HP Y

Input wattage Y

Ambient temperature (computer side) Y TC
139

Recorded in 

Data File Alarms Signal Type

Ambient temperature (test fixture side) Y TC

Smoke detector Y Smoke Detected

Setpoint Y

Bypass valve Y

Hydraulic actuator pressure Low Pressure DC Volts

Involute oil flow Y Pulses

Involute oil temp out Y

Involute oil temp in Y

Involute particles Y

Involute oil pressure at manifold Y Loss of Oil Pressure DC Volts

Involute, outside high speed/intermediate separating 

force Y SG

Involute, outside low speed/intermediate separating force Y SG

Involute, outside, low speed bearing temperature Y Excessive Temperature TC Isolated

Involute, outside, intermediate bearing temperature Y Excessive Temperature TC Isolated

Involute, outside, high speed bearing temperature Y Excessive Temperature TC Isolated

Involute, inside high speed/intermediate separating force Y SG

Involute, inside low speed/intermediate separating force Y SG

Involute, inside, low speed bearing temperature Y Excessive Temperature TC Isolated

Involute, inside, intermediate bearing temperature Y Excessive Temperature TC Isolated

Involute, inside, high speed bearing temperature Y Excessive Temperature TC Isolated

Involute reservoir temperature TC

Involute Gearbox Sump Temperature Y TC

Involute magnetic chip detector Y Particles Detected DC Volts

Involute filter pressure drop Y Excessive Filter Pressure Drop

Involute X vibration Y Excessive Vibration AC mV

Involute Y vibration Y Excessive Vibration AC mV

Involute Z vibration Y Excessive Vibration AC mV

Involute total vibration (calculated) Y Excessive Vibration calc

Involute heat exchanger oil temp in TC

Involute heat exchanger oil temp out TC

Involute particle count 4u DC Volts

Involute particle count 6u High Count DC Volts

Involute particle count 14u High Count DC Volts
140

Appendix F. Involute Gear Rating per AGMA 2001

Low Speed Gear Set
141
142
143
144
145
146
147

High Speed Gear Set
148
149
150
151
152
153
154

Appendix G. Comparison of Stress Levels at 200% Rated Load

13 October 2005

Summary of Gear Bending Strength and Surface Durability Ratings 

Reference industry standards: ANSI/AGMA 2001-E04

An analysis of all gear elements was performed to confirm that the gear tooth bending strength exceeds 

the expected surface durability under the applied test loads (200% of normal rated load). Using the gear 

industry helical gear rating standard, each involute gear set was studied to determine its safety factor with 

respect to the applied test load. Stresses for the Convoloid gear form have been calculated using an 

incremental method based upon actual tooth contact and stress allowables from the industry standard. 

Load distribution factors, dynamic factors, and allowable stresses are similar in both studies.

Results of the Stress Analysis

Involute

High Speed Gear Set Pitting Bending-Pinion Bending-Gear

Pinion Torque (lb-in) 15,750

Load Distribution Factor 1.25

Dynamic Factor 1.05

Stress Number (psi) 174,376 43,634 46,683

Allowable Stresses (psi) 225,000 65,000 65,000

Stress Cycle Factors 0.982 1.004 1.037

Safety Factors 1.267 1.496 1.443

Low Speed Gear Set Pitting Bending-Pinion Bending-Gear

Pinion Torque (lb-in) 94,500

Load Distribution Factor 1.30

Dynamic Factor 1.05

Stress Number (psi) 204,893 56,127 64,558

Allowable Stresses (psi) 225,000 65,000 65,000

Stress Cycle Factors 1.060 1.037 1.219

Safety Factors 1.164 1.200 1.227

Convoloid

The torques, load distribution factors, dynamic factors, allowable stresses, and stress cycle factors used to 

calculate the stresses are the same as those used by Mr. Don McVittie in his reports HS Inv Opt for Pitting 

AGMA and LS Inv Opt for Bending AGMA. 

High Speed Gear Set Pitting Bending-Pinion Bending-Gear

Pinion Torque (lb-in) 15,750

Stresses 145,889 32,737 34,722

Load Distribution Factor 1.25

Dynamic Factor 1.05

Factored Stresses (psi) 167,137 42,967 45,573

Allowable Stresses (psi) 225,000 65,000 65,000

Stress Cycle Factors 0.982 1.004 1.037

Factored Allowable Stresses (psi) 220,950 65,260 67,405

Safety Factors 1.322 1.519 1.479
155

Low Speed Gear Set Pitting Bending-Pinion Bending-Gear

Pinion Torque (lb-in) 94,500

Stresses 159,223 36,290 37,002

Load Distribution Factor 1.30

Dynamic Factor 1.05

Factored Stresses 186,025 49,536 50,507

Allowable Stresses 225,000 65,000 65,000

Stress Cycle Factors 1.060 1.037 1.219

Factored Allowable Stresses 238,500 67405 79,235

Safety Factors 1.282 1.361 1.569

Summary

In all instances, contact stress safety factors (pitting) are lower than bending stress safety factors. 

Therefore, the gearing should fail in pitting before a tooth breakage failure should occur. The Convoloid 

stresses are always lower than the corresponding involute stresses. The torque values used here are the 

same as those used in submitted by Mr. Don McVittie in his Test Design Review of 18 March 2004. 

These torques are slightly more than 200% of the rated torque (i.e., that required to produce 108 kW at an 

field generator speed of 1,220 rpm), but they correspond to the appropriate loads experienced by the 

gearbox after taking into account efficiency losses throughout the wind turbine system. Test speed is to be 

1,780 rpm. 
156

Appendix H. Micon 108 Gearing Life Calculation

13 October 2005

Summary of Gear Surface Durability Life

Reference industry standards: ANSI/AGMA 2001-C95

Reference study documents: Task 3.2—COMPARISON OF STRESS LEVELS AT 200% RATED 

LOAD

The equation for allowable life was derived from AGMA 2001-C95 from equations 4 and 47, and Figure 

17. Equation 27 was used to confirm the equality statement in equation 4 for the involute gearing. The 

stresses applied to each gear element were reported in the referenced document. The time of test, as 

required in the test protocol, is 200 hours. 

Results of the Pitting Life Analysis

Involute

High Speed Gear Set

Test protocol stress: 174,376 psi 

Allowable stress number: 225,000 psi 

Calculated Stress Cycle factor for pitting: 0.775004 

Life: 8,872 hours 

Low Speed Gear Set:

Test protocol stress: 204,893 psi 

Allowable stress number: 225,000 psi 

Calculated Stress Cycle factor for pitting: 0.910636 

Life: 2,985 hours 

Convoloid

High Speed Gear Set

Test protocol stress: 167,137 psi 

Allowable stress number: 225,000 psi 

Calculated Stress Cycle factor for pitting: 0.742831 

Life: 18,916 hours 

Low Speed Gear Set:

Test protocol stress: 186,025 psi 

Allowable stress number: 225,000 psi 

Calculated Stress Cycle factor for pitting: 0.826778 

Life: 16,756 hours 
157

Appendix I. Component Analysis at 240% Rated Load

TASK (3.6) 

18 January 2006 (rev. A) 

Summary of Component Safety Factors 

NREL CONTRACT #ZAM-5-33200-12 

Calculations based upon increased test procedure loads (i.e., 240% torque, 200 hours life, 2,700 rpm HS 

shaft speed). Bearing lives are catalog L1 (99% reliability). Key stresses include the 2.0 overload factor as 

represented in AGMA6001. Shaft stresses are calculated per AGMA6001.

BEARINGS

High Speed Pinion:

• Applied Power: 834 hp 

 Tapered Roller Bearings (# 32312, # 33215) 

• Speed: 2,700 rpm 

• Calc. Torque: 19,468 lb-in (applied torque is 240% operational torque) 

• Minimum Bearings L1 = 213 hours 

Intermediate Pinion:

• Applied Power: 834 hp 

 Spherical Roller Bearings (# 22319CJ) 

• Speed: 448.8 rpm 

• Calc. Torque: 117,119 lb-in (applied torque is 240% operational torque) 

• Minimum Bearings L1 = 318 hours 

Low Speed Pinion:

• Applied Power: 834 hp 

 Cylindrical Roller Bearings (# NCF 2948) 

• Speed: 100 rpm 

• Calc. Torque: 525,629 lb-in (applied torque is 240% operational torque) 

• Minimum Bearings L1 = 2,282 hours 

KEYS

Note: All key calculations based upon AGMA6001 inclusion of 2.0 overload factor. 

• Applied Torque: 19,468 lb-in

High Speed Shaft (Involute)

• Key Size: 1⁄2 in x 1⁄2 in x 3 3⁄4 in

• Tensile Yield Strength: 61,000 psi (applied torque is 240% operational torque)
158

• Minimum Shear Safety Factor = 1.03

• Minimum Compressive S.F. = 1.29 

• Applied Torque: 19,468 lb-in

High Speed Shaft (Convoloid)

• Key size: 3⁄4 in x 3⁄4 in x 3 in 

• Tensile Yield Strength: 61,000 psi (applied torque is 240% operational torque)

• Minimum Shear Safety Factor = 1.70

• Minimum Compressive Safety Factor = 2.09 

• Applied Torque: 117,119 lb-in

Intermediate Pinion

• Key Size: 5/8 in x 1.1 in x 3.15 in 

• Tensile Yield Strength: 130,000 psi (applied torque is 240% operational torque) 

• Minimum Shear Safety Factor = 1.34

• Minimum Compressive S.F. = 1.01 

• Applied Torque: 525,629 lb-in

Low Speed Shaft

• Key Size: 1 1⁄4 in x 2.2 in x 5.19 in 

• Tensile Yield Strength: 75,000 psi (applied torque is 240% operational torque)

• Minimum Shear Safety Factor = 1.35

• Minimum Compressive Safety Factor = 1.03 

SHAFT STRESS ANALYSIS—INTERMEDIATE PINION

The location of concern is the intermediate gear shoulder (0.09 radius) where a keyway bisects the 

shoulder. AGMA6001 barely addresses this case, indicating:

“Experimental verification is preferred for superposition of stress concentration factors. 

Without verification, the smaller values should be used.” 

Therefore the more severe condition has been analyzed (shoulder with radius). As with the key stress 

analysis, a 2.0 overload factor is included in the peak load stress calculation (but not for the fatigue 

factors). This results in a (raw) safety factor that is twice the indicated value, providing some latitude for 

the overlap of stress concentration factors.
159

Applied Torque: 117,119 lb-in % Alternating: 25%

Applied Moment: 75,753 lb-in % Alternating: 100% 

Applied Shear Force: 19,372 lb % Alternating: 100% 

Shaft Diameter: 3.979 in. Shoulder Dia.: 4.570 

Radius: 0.09 in Surface Finish: 63 rms 

Peak Load Safety Factor = 1.28 

Fatigue Safety Factor = 1.00

By creating a smoother radius (to 32 rms from 63 rms) the fatigue safety factor increases to 1.06. 

The AGMA standard assumes a stress cycle factor of 1.0 for all cycles above 106. Operating at a shaft 

speed of 448.8 rpm, this number of cycles is reached in 37 hours. 
160

Appendix J. Manufacturing Sensitivities of Convoloid Gearing and 

Their Effects on Rating Factors

18 November 2005 

Reference industry standards: ANSI/AGMA 2001-E04, AGMA 908-B89, AGMA 925-A03

The mesh action of the Convoloid tooth form is similar to the involute form in that both exhibit a 

conjugate action. This action creates a smooth transfer of power from one component to the other. 

Because of geometry differences between the Convoloid and the involute tooth forms, however, 

calculation methods for key factors used to rate a gear set are modified. Many rating factors are the same 

and react to manufacturing deviations similarly, but differences exist where factors are based upon tooth 

profile geometry. Section numbers refer to task numbers in the Detailed Task Summary (see Appendix C) 

3.4.1. Specific Sliding—There are two different accepted ways of calculating specific sliding—one 

commonly is used by the lubrication industry, and the other is used by the gearing industry. The gearing 

industry defines specific sliding at a point on the pinion as the sliding velocity, divided by the roll velocity 

of that point on the pinion. To maintain consistency, Genesis chose the latter method in its analysis of this 

parameter, and determined the following results.

Table J.1. Comparison of the Specific Sliding Velocities for Convoloid Versus Involute Gear Sets Used in 

the Micon 108 Wind Turbine Gearbox

High Speed Gear Set Specific Sliding

Pinion SAP Pinion O.D. Gear SAP Gear O.D.

Involute -1.4487 0.5586 -1.2656 0.5916

Convoloid -0.7979 0.5168 -1.0696 0.4438

Low Speed Gear Set Specific Sliding

Pinion SAP Pinion O.D. Gear SAP Gear O.D.

Involute -2.2828 0.6578 -1.9226 0.6954

Convoloid -1.0241 0.5702 -1.3269 0.5059

From this summary data, it can be seen that the relationship of the sliding to rolling velocities for 

Convoloid gear sets are reduced throughout the mesh. This reduced “slip” of one gear tooth with respect 

to the other is highly desirable, reducing the risk of scoring and enhancing film-thickness characteristics.

3.4.2. Absolute Sliding Velocities—Values for the sliding velocity are computed from the difference in 

rolling velocities of the pinion and gear. 
161

Table J.2. Comparison of the Absolute Sliding Velocities for Convoloid Versus Involute Gear Sets Used in 

the Micon 108 Wind Turbine Gearbox

High Speed Gear Set Absolute Sliding Velocity (in/sec)

Pinion SAP Pinion O.D. Gear SAP Gear O.D.

Involute -64.1472 106.3861 -106.3861 64.1472

Convoloid -70.0374 106.8713 -106.8713 70.0374

Low Speed Gear Set Absolute Sliding Velocity (in/sec)

Pinion SAP Pinion O.D. Gear SAP Gear O.D.

Involute -20.6614 36.9765 -36.9765 20.6614

Convoloid -24.5015 36.0520 -36.0520 24.5015

From this summary data, it can be seen that the sliding velocities are slightly greater for the Convoloid; 

however, the relationship of the sliding ratio to the rolling ratio (specific sliding) is smaller. This slight 

increase in absolute sliding velocity is thought not to be detrimental, due to the relative curvature of the 

Convoloid gearing. 

3.4.3. Film Thickness—Film thickness is directly related to the curvature of the elements in contact and 

their sliding velocities. In the traditional involute tooth form, the two mating parts have a convex 

curvature (similar to two cylinders rolling against each other). This contact exhibits a constantly changing 

radius of curvature for each element, creating a varying film thickness over the tooth profile. In gears with 

small numbers of teeth, this condition causes widely varying sliding velocities. In contrast, the Convoloid 

profile creates a situation where one mating tooth portion is convex and the complementary portion is 

concave (similar to one cylinder rolling inside another) and the two profiles have a constant relative 

curvature. Thus, the oil film thickness can be consistent throughout the profile and can exhibit a capacity 

for greater film thickness.

3.4.4. Center Distance Changes—Within normal center distance tolerances common today on modern 

CNC machinery, Convoloid gearing sensitivity to center distance changes will not affect operation or 

rating expectations. However, this tolerance must not allow the center distance to be closer than the 

calculated value. Outside this generally accepted range, however, changes will detrimentally affect both 

of these parameters.

3.4.5. Profile Tolerances—Analysis of this parameter from the involute and Convoloid perspectives does 

not appear to show any significant differences in treatment. Until proven otherwise, standard AGMA and 

ISO quality tolerances will be mirrored from the involute to the Convoloid; that is, for Class 12 involute 

gears, the Convoloid tolerance of this given parameter will be replicated to its involute counterpart.

3.4.6. Profile Shifts—Convoloid design protocols do not use profile shifting as is done commonly with 

involute designs. The pitch diameter is established in a manner similar to involute practice, and Convoloid 

curvatures and transition zone geometry are struck from that point according to specific formulae. 

3.4.7. Spacing Tolerances—Analysis of this parameter from the involute and Convoloid perspectives does 

not appear to show any significant differences in treatment. Until proven otherwise, standard AGMA and 
162

ISO quality tolerances will be mirrored from the involute to the Convoloid; that is, for Class 12 involute 

gears, the Convoloid tolerance of this given parameter will be replicated to its involute counterpart. 

3.4.8. Effects of Crown—Although Convoloid gearing can be crowned in a way very similar to involute 

gearing, the characteristic “twisted tooth” syndrome initially looks to be more pronounced with the 

Convoloid tooth form than with the involute form. “Twisted teeth” are characterized by exact leads at the 

pitch diameter with positive lead readings in the dedendum, and negative lead readings in the addendum 

on the same flank of the same tooth, or vice versa. This area is of special interest, and should be studied 

carefully going forward to better access its corrective procedures, affects on stress concentration factors 

(KHB), and other important parameters.

3.4.9. Lead Correction Factors—Analysis of this parameter from the involute and Convoloid perspectives 

does not appear to show any significant differences in treatment. Until proven otherwise, standard AGMA 

and ISO quality tolerances will be mirrored from the involute to the Convoloid, that is, for Class 12 

involute gears, the Convoloid tolerance of this given parameter will be replicated to its involute 

counterpart.

3.4.10. Gross Misalignment—Analysis of this parameter from the involute and Convoloid perspectives 

does not appear to show any significant differences in treatment. Until proven otherwise, standard AGMA 

and ISO quality tolerances will be mirrored from the involute to the Convoloid, that is, for Class 12 

involute gears, the Convoloid tolerance of this given parameter will be replicated to its involute 

counterpart.

3.4.11. Bearing Capacity—Although the radial and separating forces of Convoloid gearing closely 

approximate those in same sized involute gearing under the same load/speed spectrums, when Convoloid 

theory is applied to increase power density of a gear drive system to carry a certain load/speed spectrum, 

gear center distances and gear sizes are reduced as compared to the involute design (and maintain the 

same or lower Hertz and bending stresses), thus increasing the loads on the supporting bearing system. 

The bearing industry has recently developed processes and procedures to significantly increase the power 

density of their products. It is thought that these advancements will economically accommodate these 

anticipated increased loads, providing the industry with well-balanced gear drive systems (from a stress 

and reliability standpoint). 

3.4.12. Tangential, Thrust, and Radial Forces—Tangential and thrust forces are calculated in the same 

manner for Convoloid and involute gear sets with similar gear geometry (e.g., operating pitch diameter 

and helix angle). Radial (or separating) forces for Convoloid gears are comparable to involute gears with 

larger pressure angles.

3.4.13. Possible Constraints / Enhancements—Preliminary calculations involving the Convoloid tooth 

form reveal opportunities for reducing the size of wind turbine gearboxes by 15% to 20% via center 

distance reduction. However, limitations exist with the power density of available bearings. Efforts are 

underway to determine the availability of enhancements by various bearing manufacturers to solve this 

problem. 
163

Appendix K. Shakedown Protocol—Micon 108 in Preparation for 

Executing the Test Protocol

November 17, 2005 

In preparation for executing the test protocol, a preliminary test—or shakedown—is required to confirm 

proper operation of equipment and test instrumentation, filter the test oil to protocol cleanliness, and 

familiarize personnel with the equipment. This procedure assumes all equipment and instrumentation has 

been assembled and tested in the static state. The purpose of this process is to find any discrepancies and 

take necessary actions in support of the test protocol.

1. Familiarize test personnel with instrumentation and test protocol. Set up and start manual test log. 

2. Apply marking compound to teeth flanks—each tooth should be uniquely labeled. Per test protocol 

document of September 7, 2005. 

3. Obtain and label two 3-oz samples of new oil from drum for later testing.

4. Clean the test oil by operating the oil pumps and oil heaters for 6 hours, or until the oil cleanliness 

reaches -/15/12 and the oil temperature reaches 150° F, prior to startup. 

5. With the motor operating, but with 3% to 5% (228 lb to 380 lb-in) load, run the oil pumps until an -

/15/12 oil cleanliness is again achieved.

6. Take one 3-oz oil sample from each gearbox and perform a preliminary patch test.

7. Bring the load up to 10% torque (760.4 lb-in) and operate for 1 hour. Confirm that instrumentation 

operates properly. Continuously monitor particulates; if they exceed -/18/13, shut the system down 

and determine the reason.

8. Increase load to 50% torque (3,801.8 lb-in) and operate for 1 hour. 

9. Stop operation and inspect wear pattern. Photograph the results. 

10. Restart and slowly increase load to 100% torque (7,603.6 lb-in) and operate for 1 hour. 

11. Stop operation and inspect and photograph wear pattern. 

12. Repeat Step 8 for 150% (11405.3 lb-in) and 200% (15,207.1 lb-in) load with inspections and 

photographs performed after each test. Note any abnormalities. 

Note: The bulk oil temperature should not exceed 165° F, measured at inlet to gearbox, at any time. Nor 

should any bearing temperature exceed 190° F (an alarm will sound). If either of these conditions exists, 

stop the test run and contact the instrumentation engineer. 

If there is a shutdown due to mechanical problems during this shakedown procedure, the procedure will 

be restarted with zero loads after the problem is repaired. If there is a shutdown due to instrumentation or 

lubrication problems, then the protocol will be repeated from the last successful load step. (Repeat one 

load step.) 
164

Appendix L. Micon 108 Test Protocol

September 7, 2005 

References

• Previous draft of test memo for objectives and general description of the test 

• TGW layouts for test setup 

• DCE specs for instrumentation

DCE specs for instrumentation

1) Preliminary checks:

a) Gear accuracy checking and numbering

i) Number and mark with a vibrating pencil all of the teeth on all pinions and gears. Number 

clockwise looking upwind toward the rotor end of the gearbox.

ii) Identify and mark which flanks are the loaded flanks of the involute and Convoloid sets.

iii) Check the contact patterns of the loaded flanks of the Convoloid sets in the test stand after 

finish grinding.

iv) Document the contact with tape impressions and mark the pinion and gear teeth used so the 

same teeth can be used at assembly and during testing.

v) Chart the accuracy of the involute sets on M&M.

vi) Prior to charting lead, spray diemakers’ ink on half of face width to document the thickness of 

the ink.

b) Torsional stiffness of the involute gearbox: To estimate the appropriate torsional offset between 

the two gearboxes at no load, the involute gearbox will be pretested to 100% torque as follows.

i) Mount the gearbox on the test fixture, with low speed shaft Ringfeder coupling tightened to 

specification.

ii) Lock the low speed shaft to the test fixture with a torque arm and Ringfeder coupling.

iii) Attach a 10-in diameter 360° protractor disc to the input shaft.

iv) Torque the input shaft to 120% of nominal torque, recording torsional deflection at 5%, 15%, 

30%, 60%, 90%, and 120% of nominal torque. (Test torque will be 200% of nominal torque.)

v) Carefully reduce torque to zero, recording the deflection at 10% of nominal torque and zero 

torque.

vi) Plot the results of the static torque test and report them to Genesis. Genesis will provide an 

estimate of the preliminary no-load torsional offset between the test and slave gearboxes from 

the static torque test results.
165

c) Contact patterns and gearbox cleanliness

i) Tooth marking

ii) Paint all of teeth of both gear boxes with diemakers’ ink (Dykem red or blue) at final 

assembly. Touch up any teeth of the involute gearbox where marking is damaged in the 

torsional stiffness test.

iii) Select four teeth at 90° spacing on each gear of both gearboxes. Paint white dashes on the tips 

of those teeth, one dash at 0°, two at 90°, etc., for permanent identification. Number clockwise 

looking upwind toward the rotor end of the gearbox.

d) Preliminary test runs

i) Minimum load

ii) Torque up gearboxes to eliminate backlash and assemble couplings in high speed connecting 

shaft with gearboxes offset as specified by Genesis.

iii) The circulating/filter system should be run for a significant period of time until the oil 

cleanliness reaches -/15/12 before the gearboxes are rotated and loaded.

(a) Bring test rig up to speed and increase torque to 10% of nominal torque (5% of test 

torque) 

(b) Run at 10% torque for 2 hours minimum or until the oil particle count is reduced to 

-/15/12 per ISO 4406. 

(c) After 2 hours, stop operation, inspect, and record by digital photography the contact 

patterns of the four marked teeth on each gear of both gearboxes. Include a scale in 

photos for dimensional reference. Assign a unique file name to each photo to clearly 

identify. 

(d) Use numbers and paint marks on teeth for identification. 

(e) Repaint four teeth used for contact pattern checks.

Load steps: Repeat the procedure above at 30%, 60%, 90%, 120%, and 150% of nominal torque, with the 

same minimum run time, cleanliness requirement and recording contact patterns.

2) 200% load test

a) Re-apply marking compound.

b) Torque up gearboxes to eliminate backlash and assemble couplings in high speed connecting shaft 

with gearboxes offset as specified by Genesis.

c) Bring test rig up to speed and increase torque to 50% of nominal torque (25% of test torque). Run 

at 50% torque for 5 minutes.

d) Increase torque to 100% and run for 5 minutes.

e) Increase torque to 150%.

f) After 1 hour, stop operation, inspect, and record by digital photography the contact patterns of the 

four marked teeth on each gear of both gearboxes. This counts as the first hour of the 200-hour 

test.

g) Restart at zero torque, increasing torque gradually to 200% over one minute.

h) After 24 hours, stop operation and record contact patterns and tooth condition of the marked teeth 

and any damaged teeth. Take an oil sample at each stoppage. Conduct a patch test of the oil and 

compare to ISO cleanliness requirements. Repeat this step every 24 hours of running time until 

200 hours at 200% torque test is complete.
166

i) Sweep sump with a magnet prior to gear contact checks. Document debris on magnet with photos; 

collect and save debris.

3) Conditions which require stopping the test and inspection of the gear boxes.

a) Loss of test torque

b) Sudden temperature rise

c) Increase in noise level

d) Unusual noise (bangs, thumps, etc.)

e) Surface fatigue pitting—where macro pitting exceeds one half the face width of any teeth

f) Unusual contact patterns

4) Oil Samples

a) Two oil samples are to be taken each day from the same location from each gearbox and using the 

same procedures.

b) Oil samples must be specifically marked with subject gearbox and time of test. Any additional 

observations are to be documented.

c) Check and document oil cleanliness using patch test. Documentation is to include digital photo 

(with reference scale), additional observations, and original patch.

d) Retain one oil sample from each box for future oil tests.

i) Possible further analysis by an outside laboratory (e.g., Herguth Labs)

ii) Further tests could involve viscosity, particle analysis, and chemistry

5) Data Storage and Reports 

a) Test data must be downloaded from the data-loggers daily, and be stored in at least two different 

forms and/or locations.

b) General observations of the test are to be documented (noting elapsed time of the test) during each 

shut-down period and at random intervals during the test.

c) Test samples must be appropriately labeled with, at minimum, subject gearbox and time of test, 

and be stored in a manner that will preserve the integrity of the sample.

d) Daily reports

i) Verify daily that all collected data is reasonable and has been saved.

ii) Summarize all measured parameters.

iii) Note any significant events or changes.

iv) Include several photos of contact patterns.

v) Notes reference contact inspections.

e) Final Report

i) Purpose of the test

ii) Daily summary record of all measured parameters and observations

iii) Complete list of all contact pattern and patch test photos taken

iv) Conclusions
167

Appendix M. Test Protocol—Addendum A: 

Continuation of Test

September 7, 2005 

Description

Without having a definitive failure mode, additional testing is required. Analysis of components 

allows an increase of load to 240% of rated torque, and a potential speed increase (up to 2,700 rpm). 

1) Initial application of increased load and speed 

a) Follow standard protocol for normal startup to 200% torque (429,200 lb-in) and 1,768 rpm 

(60 Hz).

b) Allow monitoring system and lubrication to stabilize for 1 hour at minimum.

2) Increase torque to 220% of rated load (472,100 lb-in) with existing motor speed and allow oil 

levels and temperatures to stabilize (minimum of 15 minutes).

3) Increase torque to 240% of rated load (515,000 lb-in low speed shaft torque—1,000 uE on the 

load cell) with existing motor speed and allow oil levels and temperatures to stabilize (minimum 

of 15 minutes).

4) Increase speed to 2,000 rpm (67.9 Hz) and maintain torque at the 240% level and allow oil levels 

and temperatures to stabilize (minimum of 15 minutes). Additional increases in speed are not 

advised due to the high speed shaft coupling speed limitations. 

5) Planned shutdowns 

a) Scheduled shutdowns should occur Wednesday afternoon (1500 hours) and Saturday 

morning (0700), and should follow the shutdown procedure. 

b) Saturday shutdowns must be followed with securing the equipment for the weekend. 

6) Emergency shutdown. If at any time during the test cycle any loud or unusual noises are heard in 

the test equipment, bearing temperatures go above 185o, or excessive vibration of any of the 

equipment occurs, then perform an emergency shutdown and contact the test director. 

7) Inspections 

a) Inspections will be conducted the beginning of week on Monday morning (0900 hours) and 

midweek on Wednesday afternoon (1500 hours). At this time, perform detailed internal gear 

inspections of both gearboxes. 

b) Oil samples will be obtained from each gearbox each day at approximately 1500 hours 

following the established protocol.

c) Computer data will be downloaded each day, and a new data sequence started for each 

computer at approximately 1500 hours. 

d) After inspections, if the gearboxes can satisfactorily continue to run, execute the startup 

procedure and return to 2,000 rpm and 240% load. 

e) The testing cycle continues until the inspector reports a defined failure has occurred. This 

failure could be advanced pitting, tooth cracks or breakage, or some other failure mode. 
168

Appendix N. Description of Start-Up and Shutdown Procedures—

Micon 108

Start-Up Procedure

To ensure proper lubrication and prevent improper overloads, this procedure was followed for every 

start-up. 

• Main lubrication pumps are activated.

• Allow oil level to rise to desired oil-level mark. This level was established just above the low 

speed bearing race allowing the bearing rollers to dip into the oil, but below the level of the 

labyrinth oil seals. Oil level was moderated by adjusting ball valves installed in the oil drain 

hoses. Because a constant flow of oil was desired (to ensure clean oil and to remove heat 

from the gearbox), the amount of oil released through the drain had to equal the amount 

supplied. The drain valves required constant monitoring due to the variation in oil flow as the 

viscosity varied with temperature.

• Once the oil level was acceptable, the main drive motor was energized. The motor was 

controlled through a variable frequency drive that allowed for a soft start. Full speed was 

achieved over a 15-second time frame.

• With the drive motor at full speed (1,768 rpm—measured), torque gradually was added to the 

system by increasing pressure in the hydraulic cylinder. The resulting force from the cylinder 

induces a twist in the common low speed shaft, resulting in a system torque. The amount of 

force (in units of strain—measured by a strain gage installed in line with the cylinder) was 

increased to a predetermined value and required continuous monitoring until stable. The 

cylinder was located 48 in from the centerline of the low speed shaft; therefore a force of 

9,144 lb was required. (equating to a micro-strain value of 886) for the 200% load condition 

(16,250 lb.-in. torque at the high speed shaft). The force of the cylinder caused the bottom of 

the torque arm structure to raise approximately 11⁄4 in from the at-rest position. 

• Shut off oil reservoir heaters. Continue oil level monitoring. Monitor for any sounds that are 

not a part of the normal operation. 

Shutdown Procedure

The system was shut down once a day for oil sample collection and gear tooth flank inspection. 

• Retrieve 3 oil samples from each lubrication system (Convoloid and involute). One sample 

will be used immediately for oil particle analysis, one to be used for external oil analysis, and 

one sample will be saved in case further analyses are required.

• Gradually reduce the system torque until no load is applied to the system. The hydraulic 

cylinder linkage should be loose. 

• De-energize the main motor. The variable frequency drive reduces the motor speed to zero 

over a 15-second time frame.

• When the system is static, shut off the lubrication pump motors. Turn on the oil reservoir 

heaters (set to 160°F). 
169

Appendix O. Micon 108—Detail of Test

Test Setup

Two gearboxes were designed for this test. In using a “four-square” test, the two drives must have 

the same ratios. To achieve this result with gearboxes of differing geometries, a non-hunting tooth 

arrangement is required. Tooth counts for the involute were 16/96 (high speed) and 16/72 (low 

speed). Tooth counts for the Convoloid were 13/78 and 12/54. The resulting overall ratio is 27.00:1. 

The original gearboxes were designed for operation in a 108-kW (electrical power) wind turbine, 

therefore this is the basis for the torque application for the test. Once the efficiencies of the generator 

were accommodated, the reference 100% torque was 24,800 Nm (219,500 lb.-in.) at the low speed 

shaft. The main drive motor provided approximately 25 hp at 1,767 rpm. Therefore, there was a 

5.5% power loss through both gearboxes combined, equating to a 97.5% efficiency for each drive. 

See Appendix T for a detailed discussion of the efficiency. 

As stated in the body of this report, the two test gearboxes were rebuilt to like-new conditions. The 

first drive was assembled with high-quality (minimum AGMA Class 12) involute gearing. The 

second drive employed gearing with the Convoloid technology. All material used for the gearing 

was 4320 alloy steel that exceeded AGMA Class 2 material specifications, and was carburized and 

hardened prior to final tooth grind. 

Preliminary testing conducted in June 2005 revealed variations in the gear mesh contact patterns 

once torque application exceeded 100% of the rated load. Subsequent testing of the gearbox housing 

for stiffness when subjected to axial forces showed a significant deflection (0.007 in to 0.009 in at 

the bearing bores). This deflection was reduced dramatically when the cover plate was taper-pinned 

to the housing, essentially becoming part of the load-carrying structure. Following this modification, 

tooth contact pattern shifts were not a cause for concern.

Additionally, the early Convoloid test pinions showed hard contact and early micropitting around the 

transition zone. Verification of the profile of all gear elements by Brown & Sharpe (North 

Kingstown, RI) indicated excess material in this region. Subsequent profile development was 

conducted due to the tooth form error that caused micropitting during the 150% load test, a 

Convoloid a non-heat treated “dummy” pinion was reground to new data as provided by Dr. John 

Colbourne (Genesis Partners LP). This data series allowed for a greater concentration of points 

around the high-curvature transition zone. Subsequent results from B&S showed the gear profile to 

closely follow the electronic master with only minor errors. Additional modifications to the profile 

were added to cause a slight relief around the transition zone and to increase the curvature within the 

transition zone and minimize measurement errors within the zone.

Note that the cycle of grinding, shipping to B&S for verification, and return of the parts required a 

large time investment. Investigations are ongoing for a local or in-plant method for checking the 

Convoloid profile. Much promise exists with the introduction of freeform grinding and verification 

software that will address non-involute tooth profiles more readily. 

A final check of the “dummy” pinion showed a profile that was very close to the electronic master. 

Subsequent verification of the high speed and low speed test pinions were not as close as the 
170

“dummy” pinion, but the profile was thought to be within an acceptable range (Figure P.1), 

especially when the high-curvature areas in the transition zone and at the tip were removed from the 

curve fit. There was some concern that additional grinding would affect the strength of the part by 

removing too much of the heat treatment case depth. After careful analysis, the resulting case depth 

after grind was found to be acceptable. The gears were not verified for profile, as the grinding 

machinery was deemed very reliable. Tooth lead and crown were checked on the traditional gear 

measurement machines.

Upon assembly into the housings, all meshes (both involute and Convoloid) were painted with a 

marking compound to determine their no-load contact. With the exception of the Convoloid high 

speed pinion, all contacts patterns were acceptable. The Convoloid high speed pinion, when centered 

on the active flank, showed contact nearer the outside edge (nearer the bearing) (Figure P.2). This 

type of contact causes hard edge loading and premature failure of the gear teeth. Regrinding the high 

speed pinion with a 0.001-in helix modification brought the contact pattern into the center of the face 

width (Figure P.3). This pinion was not re-verified by B&S due to time constraints and the feeling 

that, by using the same profile geometry data, the profile should be acceptably accurate to the 

previous check. 

Crown Specification

Design specifications for crown on the Convoloid gears were calculated generally by AGMA 

standards, listed below. 

• For the HS Gear: End relief only (0.0004 in) starting 0.56 in from each gear edge face 

• For the LS Gear: End relief only (0.0003 in) starting 0.75 in from each gear edge face 

• For the HS Pinion: 0.0008in circular crown over a 3.5-in reference face

• For the LS Pinion: 0.0010-in circular crown over a 5.5-in reference face

These values can be refined using ISO 6336 and AGMA 927 where load intensity, tooth geometry, 

and other important factors are considered. Reading the Convoloid tooth patches, and taking into 

account the tooth twisting phenomenon observed, the calculation method will be revised for future 

Convoloid pairs. 

Tooth Twisting

Tooth twisting is a phenomenon that typically occurs when grinding both involute and non-involute 

forms with a single flank grinding wheel. It manifests itself, for example, as a positive lead error in 

the addendum of the gear, accurate lead at the pitch line, and negative lead error in the dedendum 

creating a cross lead on the same flank of the same tooth. Depending on manufacturing process and 

gear geometry, the lead errors could be reversed from those indicated above. Machine tool 

manufacturers of gear grinding machines have developed software to correct these situations, such 

that lead errors generally result in acceptable low to very low values.

Tip Relief

From observations made toward the end of the test run, tip relief on the Convoloid low speed gear 

appears to be correct because there is no micropitting in the low speed pinion root. The Convoloid 

high speed gear, however, has too much tip relief, as evidenced by use of only about 50% of the 
171

dedendum of the Convoloid high speed pinion, and only about 50% contact (in the upper half) of the 

high speed gear addendum. The methodology used to calculate tip relief in Convoloid gears must be 

studied and redefined. The test data, pictures, and measurements of the 108 test gears help to serve 

the objective, as would closer study of the AGMA/ISO practice in this critical area (see Figure P.18). 

Genesis partners developed new formulae which take into account the results and refines the end and 

tip relief calculations for future testing.

Test Observations

200% x 200-Hour Test

The presence of “smears” on the active surface of the gear teeth (easily visible in Figure O.13) 

indicates that some small particles went through the mesh. No damage could be felt due to these 

marks, and old marks tended to fade as the test progressed. Oil flowed through the heat exchangers 

after the filters, therefore it is thought that particles flushed from the exchangers could have caused 

the marks. These particles would have been soft aluminum, therefore explaining the lack of damage 

to the tooth surface (pictures and figures located in Appendix P). 

Once grinding was complete and the gearboxes fully assembled, the shakedown procedure was 

begun on November 22, 2005. After the oil was filtered to acceptable limits, the 50% and 100% 

loads were applied—with inspections following each application of load. The teeth were fully 

“blued” prior to the test, therefore the full-load contact was readily apparent. More of the bluing was 

removed as the loads were increased during the shakedown procedure. All contact patterns were 

acceptable. Shakedown was completed on November 28, (year). Contact patterns for both the 150% 

and 200% were acceptable, and the test protocol was begun immediately. 

Contact for the involute gear meshes (at 200% load) showed even contact across the face widths 

(Figure P.4, Figure P.5). Some bluing remains near the tops of the teeth due to tip relief. Very faint 

bluing is visible at the edges of the face widths alluding to a proper amount of lead crowning having 

been applied during manufacture. Note: As indicated later, this perceived 200% load actually only 

was 77% of rated load. 

Contact for the Convoloid gear meshes (at 200% load) shows full contact along the transition zone. 

However, there is a significant portion (~20%) of the tooth flank with bluing in the high speed mesh 

(Figure P.6). A high percentage of the tip (due to tip relief) is still blue in comparison with the 

involute gear set. Also, at the ends of the high speed face width, significant bluing remains (due to 

crowning). Therefore, because contact in these areas would have removed the bluing, there might 

not be full contact in the entire active tooth flank. As a consequence, a smaller area of contact is 

required to carry the load leading to higher contact stresses. The wear pattern in the low speed mesh 

covers the full face width (Figure P.7). 

There is some bluing remaining on the right side of the addendum and on the left side of the 

dedendum (as shown on the low speed gear—Figure P.8). During grinding of the Convoloid pinions, 

it was noticed that when a lead crown was applied to the tooth flank, the addendum and dedendum 

exhibited a slight relative “twist.” Thus, the lead of the addendum and dedendum were slightly 

different.

Initially, the Convoloid gearbox was noticeably noisier that the involute gearbox. This perception 

was attributed to the partial contact exhibited in the wear patterns. It was thought that the crowning 
172

and end relief were incorrect. Upon application of the “true” 200% load, however, the contact spread 

to the entire face width. Also, perceived noise from the Convoloid decreased below that perceived 

by the involute. 

As early as the first inspection after 18 hours of operation at the test load, fine micropitting began to 

appear on either side of the transition zone. The test continued as planned, with a special sensitivity 

towards closely monitoring the micropitting progression. Testing at 77% of the rated load continued 

up to the 179-hour point. Very slight micropitting was beginning to occur in the dedendum of the 

involute high speed pinion, located at the point of the beginning of tip relief of the mating gear. The 

low speed pinion showed distress where the edges of the gear contact the flank. In the Convoloid 

gearbox, micropitting progressed slowly around the transition zone of the high speed pinion, and 

began in the dedendum side of the high speed gear transition zone. Slight micropitting is visible on 

the low speed pinion. Noise for the Convoloid decreased slightly as the test progressed.

On December 6, it was discovered that the calculation for micro-strain on the Lebow torque sensor 

was incorrect. Consequently, instead of applying 200% of the rated load, the gearboxes were 

experiencing only 77%. Thus the preceding analysis applies for this much lower load. Once the 

discrepancy was found (after 179 hours of operation), the gearboxes were subjected to the proper 

150% and 200% load for 1 hour in accordance with the shakedown procedure. Inspection of the gear 

meshes did not reveal any adverse effects on the teeth. The contact pattern for the high speed gear 

set in both boxes improved with the increased load, and the noise within the Convoloid gearbox 

decreased. Both low speed meshes showed signs of distress due to edge loading. 

The gearboxes showed acceptable contact, therefore the test protocol was begun anew using the 

proper 200% load. After 21.85 hours of operation at the true 200% load (approximately 2.4 million 

cycles on the high speed pinion), micropitting on the involute high speed pinion is evident on all 

teeth (Figure P.10). Additionally, there are signs of distress due to edge contact near the center of the 

bearing span. The low speed pinion is showed signs of distress along both edges, and in the 

dedendum (Figure P.11). The high speed gear and low speed gear show minimal wear. 

The Convoloid high speed pinion showed a narrowing of the transition zone due to a slight widening 

of the micropitting (Figure P.12). Full contact is being made, as the bluing now is beginning to wear 

on the side nearest the bearing. Some distress due to edge loading was visible on the opposite side of 

the face width.

The Convoloid high speed gear is showing light micropitting around the transition zone, with the 

heaviest in the dedendum (Figure P.13). There still is a significant portion of the profile (at the tip) 

that is not making contact. This fact causes a higher than expected stress in other areas of contact.

The Convoloid low speed pinion shows micropitting advancing along the face width (Figure P.14). 

Distress is evident at both edges and at the tip. The low speed gear is showing faint signs of distress 

at along the transition zone and at the bottom of the dedendum (Figure P.15). 

After 91.25 hours of operation at the true 200% load (approximately 9.8 million cycles on the high 

speed pinion), micropitting is progressing slowly on the involute high speed and low speed pinions. 

No wear is evident on the gears. The increase in micropitting seems to correspond with an increase 
173

in noise from the involute gearbox. Figure P.16 shows a close-up of the micropitting in the high 

speed pinion. 

Slow growth of micropitting on the Convoloid high speed pinion is narrowing the width of the 

transition zone. Otherwise there is no dramatic change in the micropitting. Of special interest is that, 

during run-in, marks appeared on the tooth flanks due to debris in the oil. Most of these have not 

precipitated wear, but one mark seems to be leaving an impression on the mating flanks of the gear 

(Figure P.17). These marks later began to fade, removing concern that they would precipitate an 

unwanted failure. Noise from the Convoloid gearbox continued to decrease, and was comparable 

with the involute gearbox at this time.

Micropitting is consistently evident on the high speed gear at the dedendum side of the transition 

zone (Figure P.18). The photo in Figure P.18 also shows very well the area of no contact at the tooth 

tip due to excessive tip relief. This area represents a significant portion of the contacting flank, yet is 

not carrying any load, causing the rest of the tooth flank to see higher stresses than calculated. 

At the conclusion of testing, the gearboxes had been subjected to approximately 179 hours of the 

77% load and 235 hours of the true 200% load. Significant micropitting is evident in the involute 

high speed pinion near the root and along one edge (Figure P.19, Figure P.20) and appears to expand 

at a significant rate. This wear appears to have caused a degradation of the profile leading to a 

significant increase in noise from the involute gearbox. Conversely, the high speed gear shows little 

wear (Figure P.21). The low speed pinion shows light micropitting around the edges and at the peaks 

of grinding grooves (Figure P.22). As with the high speed gear, the low speed gear shows minimal 

wear (Figure P.23). Vibration levels within the involute gearbox increased approximately 6% during 

the time of the 200% test. 

The Convoloid gear sets show definite micropitting on all tooth surfaces around the transition zone. 

However, in contrast to the rapidly spreading wear of the involute high speed pinion, the 

micropitting seems to have retarded its advance during the last 100 hours of the test. The Convoloid 

high speed pinion shows distinct micropitting on both sides of the transition zone (Figure P.24), but 

there still is a definite area of no contact in this region, as is expected by the Convoloid geometry. 

The high speed gear shows slight micropitting wear along the dedendum side of the transition zone 

(Figure P.25). The addendum side remains free of wear. This figure also shows the large area of noncontact. 

The low speed pinion shows distinct micropitting around all edges, except in the root (Figure P.26). 

This “picture frame” phenomenon requires additional study. The low speed gear also shows faint 

signs of micropitting except at the tip (Figure P.27). Some shades of bluing remain at the tip of the 

low speed gear, but it is evident that contact occurred here. Thus, an appropriate amount of tip relief 

must have been applied. Vibration levels within the Convoloid gearbox decreased approximately 9% 

during the time of the 200% test. 

Oil samples taken from both gearboxes were extremely clean. Towards the end of the test 

(approximately 165 hours), metal particles were observed in the involute sample (although not at a 

dangerous level) and they continued to appear until the test was shut down. Also note that there was 

more entrained air in the involute samples, as compared to the Convoloid samples.
174

Micon 108 Extended Test (240% load)

(Pictures and figures located in Appendix P.) 

Having completed the testing protocol (200 hours at 200% load and 50% faster speed) during 

December, it was determined that the test results were not definitive. Extensive discussions early in 

the month resolved that the test should continue until a failure occurred. 

To reduce the length of the test, investigations were conducted to determine whether the applied load 

and the motor speed could be increased. Analysis of the limiting factors (gear tooth life, key stress, 

and shaft strength) confirmed that and increase of load to 240% of the operating torque was viable. 

Speed of the test motor was to be increased to as high as 2,700 rpm which increases the number of 

load cycles in a shorter period of time. An addendum to the test protocol was developed (Appendix 

M). For reference, 240% of the rated load is 515,000 lb-in torque versus the standard 214,500 lb-in.

Upon completion of the original test, the high speed pinion of the Convoloid gearbox was removed 

and sent to Brown and Sharpe (B&S) for inspection. Of particular interest was comparing the area of 

contact with a portion of a tooth where no contact or wear has occurred. On either side of the 

transition zone, micropitting has removed material, yet the geometry within the zone is identical to 

the untouched region. 

Testing resumed on January 23 and followed the test protocol addendum. Once the oil and bearing 

temperatures stabilized with a 200% load and 1,768 rpm, the load was increased to 220% and then to 

240%. No abnormal increases in temperature were observed and sound levels did not noticeably 

increase. The rubber cover for the constant velocity coupling nearest the Convoloid gearbox, 

however, was observed to slide away from the mounting shoulder. The test was halted and the cover 

was moved back to its proper position. After the test was restarted the coupling appeared to run 

smoothly. 

After confirming that the system was stable, motor speed was increased to 2,000 rpm. As the speed 

increased, noise from both gearboxes seemed to increase. The involute gearbox exhibited a higher 

frequency due to a higher number of teeth on its high speed pinion (16) compared to the Convoloid 

(13). Oil levels in both gearboxes did not seem to be affected by the increase in speed and bearing 

temperatures were satisfactory.

The motor speed was increased to 2,150 rpm. Evidently, the frequency of the vibrations from the 

involute high speed gear mesh were close to a system harmonic, because the sound from the involute 

gearbox increased dramatically, requiring the test personnel to wear hearing protection. The torque 

arm structure is different because the involute drive in that the motor also must be supported, and 

this structure could act as a sounding board. After the speed was increased to 2,350 rpm, the noise 

level—although at a higher frequency—dropped. Due to concerns with the coupling no additional 

speed increases were conducted.

After approximately 1 1⁄2 hours at the new test speed of 2,350 rpm (at 240% load), the rubber 

coupling cover nearest the Convoloid gearbox failed. The test immediately was shut down and 

examination of the coupling revealed a 1 1⁄2-in wide, 6-in long strip of the coupling that had been 

removed. The coupling manufacturer emphatically stated that the rubber cover was required for 
175

operation. A replacement cover was found and shipped to a repair facility in Portland, Oregon. After 

the coupling was repaired and the test equipment was reassembled, the test was resumed.

The manufacturer indicated a maximum speed of 2,000 rpm for the test coupling. Investigations into 

prior communications with the coupling manufacturer did not reveal any speed limitations for the 

coupling, although previous discussion had been focused on operation at 1,800 rpm. Due to the 

speed limitations of the coupling, the test was restarted with an operation speed of 2,000 rpm and 

240% load. 

After 70 hours of operation (302 total test hours), micropitting on the involute high speed pinion 

teeth (Figure P.28) appeared to have spread in the pinion root and along one edge (c.f. Figure P.19). 

Wear on the Convoloid teeth (Figure P.29) did not appear much changed from the end of the 

previous test (c.f. Figure P.24). 

Vibration levels in both gearboxes appear to be changing as the test progresses. The involute 

gearbox is seeing a gradual increase in vibration. Conversely, vibration in the Convoloid gearbox is 

slowly decreasing. Spectrum analysis of these vibrations reveals a substantial difference in the two 

gearboxes. High speed mesh frequencies at 2,000 rpm are 533 Hz for the involute and 433 Hz for the 

Convoloid. Figure P.30 and Figure P.31 compare the high speed mesh frequency for the involute and 

Convoloid respectively. The fundamental frequency for the high speed mesh is 19.2 dB higher with 

the involute gearbox. Additionally, an audio sound spectrum shows a very distinct difference in 

noise level between the two gearboxes (Figure P.32). The involute gearbox is 12.8 dB higher than 

the Convoloid gearbox, with the primary noise occurring at the high speed gear mesh frequency. 

Mr. Williston of Genesis investigated the speed limitations of the constant velocity (CV) joint that 

ties together the high speed shafts of the two gearboxes. Discussions with Mr. Fred Standfest of 

Universal Technical Services (UTS) who was instrumental in the design selection of the CV 

couplings revealed that the mechanical components of the coupling were viable up to 5,000 rpm. 

However, the joint itself is limited by the rubber “boot” that retains the lubricating grease. Mr. 

Standfest was comfortable with a 50-rpm increase in speed, but cautioned that any additional speed 

increases should be closely monitored—especially coupling temperature, as this directly affects the 

solidity of the boot. Subsequent testing at increased speeds revealed that the involute side coupling 

boot began to “billow” or expand at speeds as low as 2,100 rpm. To prevent a potential coupling 

failure, motor speed was reduced to 2,050 and testing progressed at this speed. 

Audible noise emanating from the test apparatus slowly rose to approximately 92 dB by mid-month 

and remained at this level through the end of the month. Vibration and noise levels continued to be 

closely monitored to provide early warning signs for impending failure. The computer monitoring 

system was designed with warning and alarm points for excessive temperature in any bearing, high 

particulate counts, or excessive vibrations. Spectrum analysis of vibrations from each gearbox 

allowed monitoring specific frequencies and harmonics of the drive system. As discussed in previous 

reports, amplitude levels for the involute high speed gear mesh are significantly higher than the 

Convoloid. 

One key vibration appeared mid-month that was not tied to any mechanical device in either gearbox. 

This vibration (68.3 Hz) was exactly two times the high speed shaft rotational frequency (34.15 Hz 

at 2,050 rpm). Upon closer inspection, audible noise relating to this vibration was thought to be 
176

coming from the involute side CV joint. This joint has not had any alteration and had the original 

rubber boot. These couplings are not designed for field maintenance, and after discussion with the 

manufacturer it was deemed that no action was necessary. On March 1, 2006, noise levels and 

unexplained fluctuations in an intermediate shaft bearing temperature in the Convoloid gearbox 

prompted the test monitor to shut down the test. Subsequent inspections revealed a failure of one of 

the intermediate bearings in the Convoloid gearbox. 

It is important to note that the Convoloid gearing did not show a progression in wear during the last 

200 hours of operation. Conversely, the involute gearbox experienced a continual progression in 

micropitting along the flanks of both pinions. The following figures show the relative wear from the 

initial contact patch, to conclusion of the 200-hour test, to the last inspection prior to the bearing 

failure.

Figure O.1. Involute high speed pinion—initial 200% loading; nearly full contact as the bluing has 

been nearly uniformly removed; hard contact is visible on the left side of the contact patch; 

paint marks on the tooth tip are used for identification to verify the same tooth was view in 

subsequent inspections
177

Figure O.2. Involute high speed pinion—235 hours at 200% load; micropitting (dull gray area) in 

dedendum and at the edge of contact

Figure O.3. Involute high speed pinion—235 hours at 200% load; micropitting (dull gray area) in 

dedendum and at the edge of contact
178

Figure O.4. Involute high speed pinion showing extensive micropitting; 763 hours total test time (89 x 

106 cycles). (235 hours at 200% load and 1,767 rpm motor speed, then the load was increased to 240% 

and speed increased to 2,050 rpm)

Wear on the mating involute high speed gear shows minimal wear in comparison with run-in. 

Figure O.5 Involute high speed gear at shakedown
179

Figure O.6. Involute high speed gear at shakedown after 763 hours; minimal wear is visible

Figure O.7. Involute low speed pinion at 200% shakedown; the contact patch reveals a consistent 

contact pattern across the contact flank; note paint marks on teeth tips for inspection purposes

Figure O.8. Involute low speed pinion after 235 hours at 200% rated load; light micropitting is visible 

on all four edges of contact with a heavier concentration in the dedendum; faint micropitting also is 

visible on the grinding peaks (faint lines that run horizontal in the active flank)
180

Figure O.9. Involute low speed pinion after 763 total test hours; extensive micropitting is visible in the 

dedendum and at the right side

Figure O.10. Close-up of micropitting on the right side of the involute low speed pinion after 702 total 

test hours
181

Figure O.11. Involute low speed gear after 200% shakedown

Figure O.12. Involute low speed gear after 763 total test hours; minimal wear is evident
182

Figure O.13. Convoloid high speed pinion after shakedown of 1 hour operation at 200% rated load; 

micropitting on either side of the transition zone appeared during approximately 160 hours of testing 

at 77% rated load; radial marks are due to aluminum particles from the heat exchanger that passed 

through the gear mesh; no impression could be felt nor was there any damage to the tooth flank

Figure O.14. Convoloid high speed pinion after 235 hours at 200% rated load

Figure O.15. Convoloid high speed pinion after 360 total test hours; micropitting around the transition 

zone is broader than after shakedown
183

Figure O.16. Convoloid high speed pinion after 538 total test hours; micropitting around the transition 

zone shows minimal progression

Figure O.17. Convoloid high speed pinion after 763 total test hours; micropitting zones on either side 

of the transition zone have broadened
184

Figures O.18. Convoloid high speed gear at 200% shakedown

Figure O.19. Convoloid high speed gear at 200% shakedown after 763 total test hours; micropitting on 

the dedendum side of the transition zone corresponds with wear in the mating pinion; note the large 

amount of bluing still visible at the tooth tips

Figure O.20. Convoloid low speed pinion after shakedown of 1 hour operation at 200% rated load; 

micropitting at the either side of the transition zone appeared during approximately 160 hours of 

testing at 77% rated load
185

Figure O.21. Convoloid low speed pinion after 763 total test hours

Figure O.22. Convoloid low speed gear after 200% shakedown
186

Figure O.23. Convoloid low speed gear after 763 total test hours; light micropitting is visible at the 

edges of the dedendum side of the transition zone; no micropitting is evident at the tip, nor is there 

much bluing

On March 1, 2006, noise levels and unexplained fluctuations in an intermediate shaft bearing 

temperature in the Convoloid gearbox prompted the test monitor to shut down the test. Significant 

damage was present on the rollers and cage (Figure O.24). Pieces of the cage and roller material 

passed through gear meshes and other bearings (Figure O.25). Significant quantities of metal were 

recovered in the gearbox sump, oil reservoir, and primary filter. A day prior to the bearing failure, 

magnet inspection of both gearbox sumps did not reveal any metal particles. After the failure, 

extensive quantities were found inside the gearbox housing. 

Analysis of temperature and vibration data taken during the final day of operation, shows a distinct 

increase in vibration levels approximately 8 hours before shutdown (Figure O.26, Figure O.27). An 

observation by the operator that was written in the test log at the time of the increased vibration 

indicates that there was “a change in the sound” of the Convoloid gearbox. However, testing 

continued until a more dramatic failure occurred.
187

Figure O.24. Failed spherical roller bearing in Convoloid gearbox after 763 total test hours; note pitted 

edges of the rollers and deformation of the cage; metallic particles are present at the bottom of the 

housing bore

A meeting was conducted at The Gear Works (TGW) in Seattle, Washington, to discuss the test 

results. In summary, testing of the existing gearboxes was halted pending analysis of the bearings 

and feasibility of correcting damage due to debris in the gear mesh. Harry Halloran of EUI 

emphasized that completion of a laboratory test showing a definitive comparison of the two gearing 

types was imperative. The amount of time in test: 179 hours at 77% of rated load; 235 hours at 200% 

load; 528 hours at 240% load; total test time: 942 hours. 

Figure O.25. Pieces of the failed bearing
188

Figure O.26. Bearing temperature chart for the failed Convoloid intermediate bearing; 

note the rise in temperature after 2:18 a.m.; the jagged “saw-tooth” shape is due to cycling of the heat 

exchanger fan

Figure O.27. Vibration levels for the Convoloid gearbox for the last day of operation; note the increase 

in vibration at approximately 1:00 a.m.

140

145

150

155

160

165

170

175

180

11:0612:0313:0013:5714:5415:5116:4817:4518:4219:3920:3621:3322:3023:27

0:241:212:183:154:125:096:067:038:008:57

Temperature

Time

CO Int Brg Temp

(Actually CI Intermediate Bearing Temperature)

0.0

0.5

1.0

1.5

2.0

2.5

3.0

11:0611:5112:3613:2114:0614:5115:3616:2117:0617:5118:3619:2120:0620:5121:3622:2123:0623:51

0:361:212:062:513:364:215:065:516:367:218:068:51

Time

CONVOLOID VIBRATION LEVELS

C Total Vibration Cx Vibration

Cy Vibration Cz Vibration
189

Appendix P. Micon 108 200% x 200-Hour Detail of Test

Figures and Photos

Figure P.1. Profile plot of Convoloid high speed pinion showing very small variations; the top of the 

pinion shows the tip relief
190

Figure P.2. Convoloid High speed gear mesh no-load contact patch prior to lead modification; gearing 

has been assembled into the test housing

Figure P.3. Convoloid high speed gear mesh no-load contact patch after lead modification
191

Figure P.4. HS Involute, 77% shakedown contact

Figure P.5. LS Involute, 77% shakedown contact

Figure P.6. HS Convoloid, 77% shakedown contact
192

Figure P.7. LS Convoloid, 77% shakedown contact

Figure P.8. LS Convoloid gear showing “twist”; note the blue area at the right side of the addendum 

and at the left side of the dedendum (very faint)

Figure P.9. HS Convoloid pinion showing micropitting along the transition zone (63.7 hours—(approx. 

6.8 million cycles) at 77% load) 
193

Figure P.10. Involute HS pinion with micropitting in the dedendum and edge distress (true 200% load 

after 21.9 hours) 

Figure P.11. Involute LS pinion showing edge distress (21.9 hours at 200% load) 

Figure P.12. Convoloid HS pinion with full face contact, showing transition zone micropitting and 

edge distress (21.9 hours at 200% load) 
194

Figure P.13. Convoloid HS gear showing slight micropitting (21.9 hours at 200% load) 

Figure P.14. Convoloid LS pinion showing micropitting (21.9 hours at 200% load) 

Figure P.15. Convoloid LS gear with faint signs of distress( 21.9 hours at true 200% load) 
195

Figure P.16. Close-up of the Involute HS pinion micropitting (91.2 hours at 200% load) 

Figure P.17. Close-up of oil-debris damage on the Convoloid HS pinion; this mark began to fade at the 

next inspection (91.2 hours at 200% load) 

Figure P.18. Micropitting in the Convoloid HS gear; note the extensive area of non-contact at the tip of 

the teeth, this is due to excessive tip relief applied during manufacturing (91.2 hours at 200% load) 
196

Figure P.19. Involute HS pinion at the conclusion of the 200% testing; note the extensive micropitting 

in the dedendum and at the edges; some micropitting is visible along the tip; this wear is progressing

(235.8 hours at 200% load) 

Figure P.20. Close-up of involute HS pinion with micropitting (235 hours at 200% load) 

Figure P.21. Involute HS gear with no wear at the conclusion of the 200% test (235 hours at 200% load) 
197

Figure P.22. Involute LS pinion at the conclusion of the 200% test showing micropitting wear around 

edges and at grinding peaks (235 hours at 200% load) 

Figure P.23. Involute LS gear at the conclusion of the 200% test showing no wear

(235 hours at 200% load) 

Figure P.24. Convoloid HS pinion at conclusion of the 200% test. Note that although there is extensive 

micropitting on both sides of the transition zone, it is not rapidly progressing (235.8 hours at 200% of 

the rated load) 
198

Figure P.25. Convoloid HS gear at the conclusion of the 200% test; micropitting is evident in the 

dedendum side of the transition zone, but the addendum side is free of wear (235 hours at 200% load) 

Figure P.26. Convoloid LS pinion at the conclusion of the 200% test; micropitting is present at all 

edges except for the bottom of the dedendum (235 hours at 200% load) 

Figure P.27. Convoloid LS gear at the conclusion of the 200% test; light micropitting is evident around 

the transition zone and at the root; good tip relief causes the wear pattern to fade away at the tip; 235 

hours at 200% load
199

Figure P.28. Involute HS pinion (235 hours at 200% load; 67 hours at 240% load) 

Figure P.29. Convoloid HS pinion (235 hours at 200% load; 67 hours at 240% load) 
200

Figure P.30. Vibration spectrum for involute gearbox as measured at the high speed bearing; value for 

the involute high speed gear mesh (533 Hz) is shown in the black box (Cursor 2); transmitted vibration 

from the Convoloid gearbox (433 Hz) is listed as Cursor 1; secondary peaks are harmonics 

Figure P.31. Vibration spectrum for Convoloid gearbox as measured at the high speed bearing; value 

for the Convoloid high speed gear mesh (433 Hz) is shown in as Cursor 1; transmitted vibration from 

the involute gearbox (533 Hz) is listed as Cursor 2
201

Figure P.32. Audio (sound) spectrum at a distance of 10 ft after approximately 300 hours of operation

Figure P.33. Involute HS pinion showing progression of micropitting (235 hours at 200% load; 

355 hours at 240% load) 
202

Figure P.34. Convoloid HS pinion showing minimal progression of micropitting (235 hours at 200% 

load; 355 hours at 240% load) vertical streaks are from aluminum particles (from the heat exchanger)

Figure P.35. Involute LS pinion showing substantial micropitting at the right edge and in the 

dedendum (235 hours at 200% load; 355 hours at 240% load) 

Figure P.36. Convoloid LS pinion also shows minimal progression of micropitting (235 hours at 200% 

load; 355 hours at 240% load) 
203

Figure P.37. Involute HS pinion at the end of test; wear due to micropitting is extensive along one side 

(235 hours at 200% load; 528 hours at 240% load) 

Figure P.38. Convoloid HS pinion at the end of test; wear due to micropitting is expanded slightly 

during the test, but no progression occurred on the edges (235 hours at 200% load;

528 hours at 240% load) 
204

Figure P.39. Involute HS gear at the end of test; no wear is apparent (235 hours at 200% load; 

528 hours at 240% load) 

Figure P.40. Convoloid HS gear at the end of test; some wear due to micropitting is visible on the 

dedendum side of the transition zone; a large section of the addendum still shows marking dye 

indicating too much tip relief in the profile (235 hours at 200% load; 528 hours at 240% load) 
205

Figure P.41. Involute LS pinion at the end of test; similar to the HS pinion, micropitting wear is 

concentrated along one side and in the dedendum (235 hours at 200% load; 528 hours at 240% load) 

Figure P.42. Convoloid LS pinion at the end of test; wear due to micropitting is brackets the operating 

surface except for in the dedendum due to proper tip relief on the mating gear

(235 hours at 200% load; 528 hours at 240% load) 
206

Figure P.43. Involute LS gear at the end of test; no wear is visible (235 hours at 200% load;

528 hours at 240% load) 

Figure P.44. Convoloid LS gear at the end of test; slight wear due to micropitting is visible in the 

dedendum (235 hours at 200% load; 528 hours at 240% load) 
207

Appendix Q. Micon 108 Test Recommendations

Gear Engineers, Inc. P.O. Box 70327 Seattle, WA 98127 (206) 783-3919; gearengr@mcvittie.com 

Memo Page 1 of 2 2004 September 22 

Rev. A 2004 September 23 

Genesis, LLP

Gear geometry proposal 

To: Barney Berlinger

 

I’ve reviewed the recent correspondence and calculations for gear geometry of the test involute gears 

and offer the following recommendations. 

Objectives of test:

• Compare performance of involute and Convoloid gear teeth in a typical small wind turbine 

application; 

• Surface fatigue damage; 

• Efficiency; and

• Avoid failure by tooth bending. 

Rating calculations were made using the AGMA gear rating software, ver. 2.2, which provides 

comparative ratings according to AGMA 2001 and ISO 6336. 

AGMA 6006, which governs wind turbine gear boxes, requires minimum ISO safety factors of 

1.25 in pitting and 1.56 in bending at equivalent load (nominal rating times application factor) for 

this gearing. 

The object of the 200% torque test is to run the gears at “endurance“ load where the torque is equal 

to nominal load times application factor and safety factor (safety factor squared for pitting) to 

demonstrate that this required rated capacity has been achieved.

Any calculated ISO safety factor at 200% torque over 1.0 indicates that the gears are expected to 

survive the 200% torque test. Any calculated AGMA safety factor at 100% torque times application 

factor over 1.0 indicates that the gears are expected to survive the 200% torque 200 hour test, 

because AGMA ratings have a built-in 200% overload factor to allow a limited number of electric 

motor starts at 200% of nominal torque. The load distribution factor KHß was taken as 1.30 per LVR 

calculations. The actual dimensions of available hobs from TGW were used. It is essential that no 

hob substitutions be made during manufacture. 

Grind stock allowances were adjusted to minimize the undercut at the root fillet due to hob 

protuberance. It is essential that this be closely monitored during manufacture because the hobs 

chosen are general-purpose tools and not the ideal tools for pinions with small numbers of teeth. 
208

High speed gears:

The gear geometry has been optimized for pitting resistance as in a typical small wind turbine. This 

results in the following safety factors (allowable stress / calculated stress) at the test load of 492 HP 

(366 kW) at 1,830 rpm and 200-hr test duration. 

• ISO pitting: 1.15

• AGMA pitting: 1.22 

• ISO bending: 1.95 (limited by gear wheel) 

• AGMA bending:1.34 (limited by gear wheel) 

If the mesh geometry were optimized for bending, the AGMA rated lives in pitting would be slightly 

reduced and rated bending lives would be slightly increased. In my opinion this would jeopardize the 

acceptance of the test results, because it might be alleged that the design was biased to favor a pitting 

failure.

Typical wind turbine gear mesh geometry is optimized for pitting and micropitting resistance, 

because that is by far the most common failure mode in actual turbines and in turbine gearbox 

overload acceptance tests.

Low speed gears:

The gear geometry has been optimized for bending resistance to reduce the risk of an unwanted 

bending failure. The optimization was done by adjusting the profile shift factors of pinion and gear 

without changing the helix angle, pitch, or number of teeth. This results in the following safety 

factors (allowable stress / calculated stress) at the test load of 492 HP (366 kW) at 305 rpm and 

200-hr test duration:

• ISO pitting: 1.15

• AGMA pitting: 1.12 

• ISO bending: 1.80 

• AGMA bending:1.12 

If the mesh geometry were optimized for pitting, the AGMA rated lives in pitting would be slightly 

increased and rated bending lives would be slightly decreased. I believe that that the optimization for 

bending can be justified, because there are good reasons to avoid a bending failure. 

The calculation details are attached as separate files.

Don McVittie
209

Appendix R. Micon 108 200% x 200-Hour Test Vibration Analysis

Gearbox noise is directly related to the various vibrations generated within. From the different gear 

meshes to individual bearing components, all vibrations are related to the shaft speed. Using a 

detailed analysis and a spectrum analyzer with filtering, it is possible to monitor a system for early 

wear. Each peak within the frequency band points to a different excitation, and thus a different 

component, within the gearbox. Proper analysis requires monitoring to establish baseline values. 

Subsequent readings can be compared to the baseline to determine which frequencies are changing. 

The primary focus of the vibration data readings that were collected was to monitor the changes in 

vibration as the test progressed. Detailed analysis using a spectrum analyzer was conducted late in 

the test. The specific vibration values differ greatly from the Convoloid to the involute. However, 

this is not deemed significant in that there are too many variables to determine how potential natural 

frequencies might affect the vibrations. The locations of the sensors also have an affect on the 

magnitudes. 

What is interesting is that although the involute vibration remained steady, vibrations within the 

Convoloid gearbox reduced as the test progressed. Figure R.1 shows the vibration values for the first 

full day of testing at 200% load. Compare this with Figure R.2 which was collected the last full day 

of testing. The average Convoloid value fell from 1.685 to 1.520, a 9.8% decrease. The variations in 

the total vibration values are due to cycling of the lubricant cooling fans. Vibration readings 

dramatically increased when the fans ran (especially for the Convoloid).

Spectrum analysis of the two gearboxes shows distinct spikes at the high speed gear mesh 

frequencies (385 Hz for the Convoloid and 471 Hz for the involute) (see Figure R.3, Figure R.4). 

Additionally, the harmonics of these fundamental frequencies are very evident. In fact, the 

harmonics tend to show higher amplitudes than the fundamental. A summary of Mr. Meredith’s 

notes which detail frequencies and amplitude are summarized below. Without having a baseline for 

comparison, no further analysis is possible. 

• High speed shaft frequency (1768 rpm / 60) = 29.47 Hz 

• Hydraulic dither frequency = 263.75 Hz 

Convoloid: High speed gear mesh frequency (29.47 * 13 teeth) = 383 Hz. 

Table R.1. Convoloid Frequency Spikes and Their Amplitudes 

Vertical Deflection Horizontal Deflection

Frequency (Hz) Amplitude (G.O-P) Frequency (Hz) Amplitude (G.O-P)

383 67 x 10-3 117.5 410 x 10-3

471 65 x 10-3 765.75 500 x 10-3

766 200 x 10-3 1150 243 x 10-3

1150 126 x 10-3 1530 800 x 10-3

1530 315 x 10-3 1915 315 x 10-3
210

No frequency spikes relate to the dither frequency, therefore vibrations from the loading mechanism 

are not significant. 

Involute

High speed gear mesh frequency (29.47 * 16 teeth) = 471 Hz 

Table R.2. Involute Frequency Spikes and Their Amplitudes 

(a frequency spike also occurred at 88.5 which is a harmonic of the high speed shaft speed) 

Vertical Deflection Horizontal Deflection

Frequency (Hz) Amplitude (G.O-P) Frequency (Hz) Amplitude (G.O-P)

470 364 x 10-3 470 176 x 10-3

940 389 x 10-3 940 227 x 10-3

1410 397 x 10-3 1410 330 x 10-3

1880 225 x 10-3

Motor

High speed shaft frequency = 29.47 Hz 

Table R.3. Motor Frequency Spikes and Their Amplitudes 

(a frequency spike parallel to the shaft occurred at 1410-1414 Hz

with an amplitude of 302 x 10-3 G.O-P) 

Vertical Deflection Horizontal Deflection

Frequency (Hz) Amplitude (G.O-P) Frequency (Hz) Amplitude (G.O-P)

29.5 104 x 10-3 470 280 x 10-3

78.5 124 x 10-3

88.5 46.1 x 10-3

118 56.5 x 10-3

157 52 x 10-3

470 101 x 10-3

Conclusions

At the primary frequencies of the high speed gear meshes (383 Hz for Convoloid, 471 Hz for the 

involute) the involute definitely shows higher amplitudes. The involute frequency is great enough to 

show in the Convoloid measurement. Subsequent spikes at various harmonics could be due to 

natural frequencies of the test structure and gearboxes. Due to cessation of testing, no additional 

analysis (natural frequency determinations) is available.
211

Figure R.1. Gearbox total vibration at the beginning of the test—200% load

Figure R.2. Gearbox total vibration at the end of the test—200% load

0.0000

0.5000

1.0000

1.5000

2.0000

2.5000

10:5011:5312:5613:5915:0216:0517:0818:1119:1420:1721:2022:2323:26

0:291:322:353:384:415:446:477:508:539:56

Total Vibration

Beginning of 200% Loading

Convoloid Involute

0.00000

0.50000

1.00000

1.50000

2.00000

2.50000

10:4711:5913:1114:2315:3516:4717:5919:1120:2321:3522:4723:59

1:112:233:354:475:597:118:239:35

10:4711:5913:11

TOTAL VIBRATION

(FINAL DAY OF TEST)

Convoloid Involute
212

Figure R.3. Vibration spectrum of the Convoloid gearbox; cursor is located at the HS pinion 

fundamental frequency (385 Hz)

Figure R.4. Vibration spectrum of the involute gearbox; cursor is located at the HS pinion fundamental 

frequency (470 Hz)
213

Appendix S. Test Observations—Richard L. Meredith, D.C. Energy

Observations From Genesis

200% Rated Load 235-Hour Test Data

December 23, 2005 

Richard L. Meredith, D C Energy, Bothell, WA 

1. The hottest bearing in the Convoloid gearbox was the outside low speed which in the worst 

case was 1°F above its sump with a maximum of 167°F and 166°F respectively.

2. The hottest bearing in the involute gearbox was the inside intermediate bearing which never 

exceeded 163°F (which was also only 1°F above that sump temperature). 

3. Both reservoirs were controlled to a maximum of 161°F and a minimum of 145°F.

4. Over the 235-hour course of the test, the vibration level of the involute gearbox tended to 

increase slightly. Less than 6% increase was noted.

5. Over the same period, the Convoloid gearbox vibration level tended to decrease slightly. 

Around a 9% change downward was noted. 

6. As calculated above, the total losses for both gearboxes were between 5.26% and 5.52%. This 

includes the bearing losses and the oil churning losses. 

7. Throughout the test, the ISO cleanliness level for 6u particles in the Convoloid gearbox was 7, 

according to the online particle counter. 

8. Throughout the test, the ISO cleanliness level for 6u particles in the Involute gearbox was 11, 

according to the online particle counter. 

9. The foregoing two points were confirmed by daily patch tests for each gearbox. Patches were 

made into slides and have been preserved. 

10. There was never a “hot” bearing or any unusual noises from either gearbox during the 235-

hour run. 

11. The oil samples that were collected every day from each gearbox indicated that there was more 

entrained air in the oil from the involute box. No concrete reason for this could be determined. 

12. Due to the way that the prime mover motor was mounted in relation to the involute gearbox, 

there was approximately 700 FPM air blowing on the outside of the involute gearbox; partially 

accounting for the fact that the heat exchanger fan on the involute box less frequently than that 

of the Convoloid box. Additionally, due to the design of the heat exchanger used on the 

Convoloid box, there was more cooling due to normal flow through the core even when the fan 

was not running. The temperature drop across this heat exchanger, with normal flow, was 3° to 

5° F as compared to 1°F for the Convoloid box. 

13. Daily physical inspections did not reveal any impending catastrophic failures in either gearbox. 

At the completion of the 235.8-hour test run, both gearboxes were running normally in all 

observable respects with nothing noted that would preclude further test running. 

Richard L. Meredith

D C Energy

Test Instrumentation Supervisor.
214

Appendix T. Derivation of Gearbox Efficiencies

Overall System Efficiency

In a four-square test setup, the only power required by the system is that required to rotate the shafts. 

The amount of power supplied by the motor is equal to the losses within the system. Therefore, by 

determining the motor power supplied to the system, the losses can be found. Two methods were 

used to ascertain the power loss: amount of variation in shaft speed of the motor—no-load versus 

load; and electrical power consumption. This discussion is taken from a fax sent by Mr. Meredith to 

Mr. Williston on December 21, 2005.

Input Power Determination

Motor power based upon motor speed: 

• Test conditions: Motor speed at 200% gearbox load, measured with a calibrated strobe 

tachometer = 1,768 rpm 

• No-load speed = 1,775 rpm (test system attached, yet no loading to the system) 

• Open shaft motor speed = 1,778 rpm (no equipment attached to the motor shaft) 

Conclusion

The output frequency of the variable frequency drive (VFD) for the motor is less than 60 Hz. 

Because 60 Hz would result in ~1,800 rpm shaft speed, the actual frequency provided by the VFD is 

as follows.

( ) 1778 60 ⋅

1800 = 59.27 HZ

The motor nameplate shows 1,775 rpm for the 60 HP delivered power. Therefore, there is a full load 

(FL) slip in motor speed of 25 rpm. The full load power factor is given as 0.90. The full load motor 

line current is given as 67A.

Measured slip at the operating frequency if 1,778 - 1,768 = 10 rpm. 

10rpm

25rpm ⋅60HP = 24 HP test load

Or 24/60 = 40% of motor horsepower. 

Motor power by electric input: 

W E I PF = ⋅⋅ ⋅ 3

Where: W = electrical power (watts); 

 E = voltage (rms) = 477 volts rms; 

 I = current (amps) = 35.5 A; and 

 PF = Power Factor (a typical power factor at 40% load) = 0.64. 
215

W = ( ) ( ) ( ) ( ) 477 35.5 .64 1.732 18.77 = kW

Converting to horsepower:

1 18.77 * 25.16

.746

hp kW hp kW

    =  

There is a fairly good agreement between these two methods of arriving at the input power (losses). 

Circulating Power

* (439000 )*65.48 456.1

63025 63025

torque speed lb in rpm HP hp    −  =    =  =    

Efficiency

Efficiency % = (1—loss) * 100% 

Overall losses for the system:

• Motor Speed method: (24 hp / 456.1 hp) = 0.0526 

• Electrical Power method: (25.16 hp / 456.1 hp) = 0.0552 

The resultant system efficiency is 94.74% and 94.48% respectively, yielding individual gearbox 

efficiencies of 97.4% and 97.2%. 

Gearing Efficiency

The following discussion references AGMA ISO 14179-1: Gear Reducers—Thermal Capacity 

Based on ISO/TR 14179-1. 

Every component within a gearbox generates a certain amount of heat during operation. This heat is 

converted from mechanical power due to inefficiencies in the component. An approximation of these 

losses is given in AGMA ISO 14179-1. By removing the losses due to components other than the 

gearing, an approximate gearing efficiency can be determined. The additional step of removing the 

gear churning losses and the load-based losses of the involute gear set could provide an approximate 

idea of the Convoloid load-based losses. Thus, the efficiency of the Convoloid gear set can be found. 

Non-gearing losses consist of heat generated in the bearings and by contacting oil seals. Two types 

of loads must be calculated: Load-based in which the amount of heat generated is affected by the 

amount of load applied, and non-load based in which the generated heat is strictly due to friction 

and/or oil churning losses based on rotational speed. Load-based losses are calculated with 456.1 hp. 

Table T.1 shows the heat (in horsepower) generated in the involute gearbox. 

Table T.1. Involute Non–Gear Inefficiency Losses

Component Type of Load Shaft Speed (rpm) Power Loss (hp)
216

High Speed (2 seals) Non-load 1768.0 0.072

Low Speed (2 seals) Non-load 65.5 0 (non-contacting seals)

HS Bearings (2) Non-load (churning loss) 1768.0 0.031

Load-based 0.771

INT Bearings (2) Non-load (churning loss) 294.7 0.029

Load-based 0.691

LS Bearings (2) Non-load (churning loss) 65.5 0.028

Load-based 0.500

TOTAL: 2.12 hp

Table T.2. Convoloid Non–Gear Inefficiency Losses

Component Type of Load Shaft Speed (rpm) Power Loss (hp)

High Speed (1 seals) Non-load 1768.0 0.036

Low Speed (2 seals) Non-load 65.5 0.0 (non-contacting seals)

HS Bearings (2) Non-load (churning loss) 1768.0 0.031

Load-based 0.811

INT Bearings (2) Non-load (churning loss) 294.7 0.029

Load-based 0.889

LS Bearings (2) Non-load (churning loss) 65.5 0.028

Load-based 0.630

TOTAL: 2.49 hp

Using the electrical power inefficiency, the total non-gearing component losses equate to 4.61 hp. 

Therefore the total losses of only the gearing in the test becomes: 25.16 hp - 4.61 hp = 20.55 hp; 

giving a gearing total inefficiency of: (20.55 hp / 456.1 hp) = 0.0450. 

The resultant efficiency for the gearing is 95.5 % for the system, or 97.8% for each gearbox yielding 

a per–gear mesh efficiency estimate of 98.9% for both the involute and Convoloid pairs. 
217

Appendix U. Convoloid Gear Separating Force Qualification

Comparison of Convoloid Gear Performance To 

AGMA Rating Formulae and Factors—Micon 108 

February 6, 2007

A key aspect of designing a gear drive system is to provide sufficient support for the gears. 

This foundation takes the form of bearings and the housing supporting the bearings. 

Calculation of forces induced by the torsional loads applied to the gears is necessary to 

properly size the bearings and to confirm the strength of the housing. When torque is applied 

to helical gearing, reactive forces occur in all three planes. For Convoloid gearing, two of 

these forces, tangential and axial (or thrust), directly follow existing load theory. However, 

due to the difference in profile of the Convoloid tooth form to the involute, the third load, 

the separating force, must be calculated in minute increments with the resultant being the 

sum of the forces. Although there is much confidence in these calculations, it is prudent to 

qualify these forces.

During the operational test, two Micon 108 gearboxes (one with involute gearing and one 

with Convoloid gearing) were operated in a “four-square” or back to back type test. The 

housings of the gearboxes were instrumented with, among other things, strain gages to help 

understand how the housing deflected when loaded. Three different test loads were applied 

for a significant period of time during the test giving three distinct groups of data. 

Strain gages were permanently mounted on both sides of the housing and located between 

the high speed (HS) and intermediate (INT) bores; and between the INT and low speed (LS) 

bores. These gages, arranged in a Full Poisson Bridge (Wheatstone Bridge), allow for a high 

degree of sensitivity (2.6 times the sensitivity of a quarter bridge) and compensate for 

temperature variations. To have a point of reference in which to reduce the copious amount 

of data from the operational test, values obtained from the strain gages must be compared to 

strain values obtained when the housing is subjected to known loads. 

To develop a load / strain cross reference, a static load test was conducted on the housings. 

In addition to the Convoloid housing, the involute housing was also tested to qualify the 

method. If the resulting strains can be correlated to the loads for involute gearing, then the 

method should reasonably predict the loads for the Convoloid gearbox. These loads then 

could be compared to the calculated loads. The existing strain gages were able to be used, 

preserving the consistency with the operational test.

Static Load Test Equipment

In an attempt to duplicate how the loading of internal gear loads impact the housing (through 

the bearings), machined plugs were fitted into the bearing bores (Figure U.1). Steel tubing is 

fitted through these plugs and hydraulic cylinders are used to apply separating forces 

between the bores. The cylinders are placed within the housing in the approximate positions 

of the gear mesh they are duplicating (Figure U.2). 
218

The amount of force produced by the hydraulic cylinders is the product of hydraulic 

pressure and cylinder surface area. The HS cylinder has an area of 3.14 in2

, and the area of 

the LS cylinder is 7.22 in2. By applying a given amount of pressure (to produce a known 

load) to both cylinders, the resulting strain can be gathered. This data should give a good 

baseline that can be used for analysis of values obtained during testing. 

To preserve the viability of the test, none of the strain gages have been removed or altered 

from the original test configuration. Comparison of the strain values gathered during the 

operational test and these static values allows for a direct determination of the actual loads 

that were imposed on the housing during the test. It is important to compare the actual loads 

with theoretical calculations in order to validate or refine those calculations. Also, 

comparison of the static and dynamic test values obtained from the involute gearbox should 

give insight into potential measurement error and unknown variations the test could have. 

Figure U.1. Static load test equipment set-up 

As can be seen from the photos, access through the bearing bores to the cylinders was not 

possible. Therefore, plumbing for the hydraulic cylinders required that the coverplate not be 

installed. Early in the project, the housings were found to be unacceptably flexible when 

subjected to thrust loads (induced when setting the tapered roller bearings). The faces of the 

housing literally “bowed” out, preventing the bearings from having the proper setting. To 

counter this deflection, the coverplate was pinned in four places with removable taper pins. 

Machined

Plug

Steel

Tubing
219

Subsequent measurements showed that the deflection at the high speed bearing bore had 

been reduced from 0.009 in (at a load equal to that produced by the 200% test load) to 

0.0015in. As this test was focused on qualifying the separating loads (the forces that try to 

push the two gears in mesh away from each other) and not thrust, it was felt that any support 

of the cover would be minimal and could be neglected. 

Test Results

The amount of loading applied is related to the forces produced by the torques of the 

operational test. At the end of the test, a torque equal to 240% of the rated load was applied. 

This equates to an approximate separating force of 5,200 lb between the HS and INT bores, 

and a force of 16,800 lb between the INT and LS bores. The range of forces to be applied by 

the hydraulic cylinders began with no loading (to establish the strain gage offset) and 

increased (in 500-psi increments) to 3,000 psi (~9,400 lb) on the HS cylinder and 2,500 psi 

(~18,050 lb) on the LS cylinder. The amount of pressure on the LS cylinder was limited to 

prevent an excessive amount of load on the housing. 

Hydraulic pressure was initially applied separately to the cylinders to qualify the operation 

of the strain gages. As expected, strain appeared to follow increases in the applied pressure 

(and thus load) in a linear manner. Values from the Convoloid housing were higher than 

with the involute housing (Figure U.3, Figure U.4). 

Figure U.2. Close-up of a hydraulic cylinder in position to apply a separating load
220

Figure U.3. Representative strain/force diagram for the involute housing—HS outside position

Figure U.4. Representative strain/force diagram for the Convoloid housing—HS inside 

position

y = 2E-09x + 3E-07

R2 = 0.9952

0.000E+00

2.000E-06

4.000E-06

6.000E-06

8.000E-06

1.000E-05

1.200E-05

1.400E-05

1.600E-05

1.800E-05

0 1000 2000 3000 4000 5000 6000 7000 8000 9000 10000

Strain

Pounds Force

Gage Pair A HS/INT

Chart 1

y = 2E-09x

R2 = 0.9977

0.000E+00

2.000E-06

4.000E-06

6.000E-06

8.000E-06

1.000E-05

1.200E-05

1.400E-05

1.600E-05

1.800E-05

2.000E-05

0 1000 2000 3000 4000 5000 6000 7000 8000 9000 10000

Strain

pounds

Conv Gage Pair A HS/ Int

Chart 21
221

When pressure (load) was applied to both cylinders, strain readings were inconsistent with 

the amount of applied load. Figure U.5 shows the same strain gage as Figure U.3 being 

loaded with both cylinders. Strain is predictable until the HS cylinder is increased without 

increasing the LS.

For the strain gages between the LS and INT bores, values vary drastically. In Figure U.6, as 

pressures were increased, strain readings for one of the LS strain gages show a compressive 

load under low load conditions. The strains then reverse and show the expected tensile 

relationship until the relationship between the two hydraulic cylinders varies (when the HS 

pressure increases to 3,000 psi and the LS pressure remains constant at 2,500 psi).

Multiple trials were conducted to ensure that the data variations were not due to damage to 

the housings and/or the strain gages. The tests were repeatable, however, including the 

pressure/strain relationship of one of the Convoloid LS gages (Figure U.7) where the strain 

went to zero at the maximum loaded condition. 

Conclusions

Although the forces applied to the housings were intended to span the range of loads 

developed by the operational test, there was some concern that the housings might have been 

overloaded during the static test. Review of the strain readings from both tests reveals that 

the amounts of strain realized from the static test were much lower (by nearly an order of 

magnitude in some cases) than those seen in the operational test. Therefore it is assumed that 

the housings were not deformed excessively with the static test.

Because the loads applied by the hydraulic cylinders were in the range of the calculated 

separating loads, the additional strain appearing in the operational test can only be explained 

by other gear forces (tangential load and axial, or thrust, load). An example of the disparity 

between the static test and the operational test can be seen by comparing strain values in the 

range of the final test load (240% of rated torque) which can be approximated by the 

hydraulic loads at 2,000 psi. 

During the operational test, strain between the HS and INT bearings (outside) was 20.2 

micro-strain (E-06) and during the static test a strain of 17.2 was realized—well within the 

expected range. For the other side of the housing, the operational strain was 77.3 compared 

to a static test value of 5.9. This disparity can only be explained by the presence of 

substantial thrust induced by the gearing of the operational test. 

Similar comparisons occur between the INT and LS bores where the outside strain values 

for the operational test were 136 compared to 4.4 on the static test. The strain values for the 

inside position were much closer with 35 on the operational test and 27 on the static test. 

It appears that the amount of housing strain must be greatly affected by the thrust component 

of the gear loads. This is a force that was not modeled nor induced by the hydraulic 

cylinders. Additional forces (e.g., tangential loads and bearing setting forces) also can affect 

strain readings. 
222

Due to the great disparity in strain readings, an appropriate comparison with values from the 

operational test cannot be made, and the nonlinear relationship of the curves precludes any 

extrapolation. To be truly meaningful, the calibration should have been carried out with a 

good bearing set and an operational set of gears in their respective housings with torque 

arms attached to the input and output shafts. This method would have provided the thrust 

component to the strain gages. 

This test and the findings thereof have emphasized to Genesis Partners personnel and their 

associated subcontractors and consultants the need for more thorough analysis and planning 

prior to conducting similar tests.

Figure U.5. HS (outside) strain gage reading under combined load—involute 

0.000E+00

2.000E-06

4.000E-06

6.000E-06

8.000E-06

1.000E-05

1.200E-05

1.400E-05

1.600E-05

1.800E-05

0 500 1000 1500 2000 2500 3000 3500

Strain

PSI Hydraulic Pressure

Gage Pair A HS/Int & Int/LS

Chart 13

Int/LS PSI limited to 2500
223

Figure U.6. LS (inside) strain gage reading under combined loading—involute

Figure U.7. LS (outside) strain gage reading under combined load—Convoloid 

-1.000E-06

0.000E+00

1.000E-06

2.000E-06

3.000E-06

4.000E-06

5.000E-06

0 500 1000 1500 2000 2500 3000 3500

Strain

PSI Hyd Press

Conv Gage Pair D HS/Int & Int/LS

Chart 36

-3.000E-06

-2.000E-06

-1.000E-06

0.000E+00

1.000E-06

2.000E-06

3.000E-06

4.000E-06

5.000E-06

0 500 1000 1500 2000 2500 3000 3500

Strain

PSI hydraulic pressure

Gage Pair D HS/Int & Int/LS

Chart 16

Int/LS PSI limited to 2500
224

Appendix V. Tutorial on Convoloid Gearing Technology

Description

Convoloid represents a gear technology intellectual portfolio (IP) encompassing the design, 

stress analysis, manufacture, and inspection of gear teeth of a totally new design. The 

involute curve, first proposed for use in gearing in 1754, has been used almost exclusively 

for the production of geared power transmissions. Continued analysis and field experience 

has refined this tooth form to a high degree. The relatively recent proliferation of computers 

has made possible the rapid optimization of most factors involved to markedly increase the 

power density of gear drive systems. 

Trademark

The trademark Convoloid is owned by Genesis Partners LP. 

Tooth Form

The Convoloid tooth form is computer generated and has the following primary 

characteristics.

• Convex/concave contact of meshing tooth surfaces 

• Conjugacy 

• Optimized relative curvatures

• Carefully selected curvature values

An example of the Convoloid tooth form is given in Figure V.1. 

Architectures

Convoloid gear pairs primarily are intended for parallel axis helical gear architectures. 

Included in these architectures are straightforward helical-parallel axis pairs either speed 

increasing or speed decreasing, planetaries including internals and other epicyclic 

arrangements. The latter arrangements are especially well suited to the more compact gear 

system architectures of wind turbine gearboxes.

Advantages

Advantages over involutes include: 

• Considerably lower surface durability stresses,

• Considerably lower tooth bending stresses, 

• Enhanced entraining velocities and other tribological characteristics, and

• Manufacturing and inspection utilize the existing capital asset infrastructure.
225

Patents

Basic patents have been filed and issued in the United States and many other countries 

around the world. Additional related patent applications have been submitted and others will 

be submitted on a continuing basis. 

Figure V.1. Graphical display of the Convoloid tooth form
F1146-E(10/2008)

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PLEASE DO NOT RETURN YOUR FORM TO THE ABOVE ORGANIZATION.

1. REPORT DATE (DD-MM-YYYY)

August 2010

2. REPORT TYPE

Subcontract Report

3. DATES COVERED (From - To)

June 2005 - October 2008

4. TITLE AND SUBTITLE

Low Wind Speed Turbine Developments in Convoloid Gearing:

Final Technical Report, June 2005 - October 2008

5a. CONTRACT NUMBER

DE-AC36-08-GO28308

5b. GRANT NUMBER

5c. PROGRAM ELEMENT NUMBER

6. AUTHOR(S)

Genesis Partners, LP

5d. PROJECT NUMBER

NREL/SR-500-45103

5e. TASK NUMBER

WE10.1161

5f. WORK UNIT NUMBER

7. PERFORMING ORGANIZATION NAME(S) AND ADDRESS(ES)

Genesis Partners, LP

209 Progress Drive

Montgomeryville, PA 18936

8. PERFORMING ORGANIZATION

REPORT NUMBER

ZAM-3-33200-12

9. SPONSORING/MONITORING AGENCY NAME(S) AND ADDRESS(ES)

National Renewable Energy Laboratory

1617 Cole Blvd.

Golden, CO 80401-3393

10. SPONSOR/MONITOR'S ACRONYM(S)

NREL

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AGENCY REPORT NUMBER

NREL/SR-500-45103

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National Technical Information Service

U.S. Department of Commerce

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13. SUPPLEMENTARY NOTES

NREL Technical Monitor: Scott Schreck

14. ABSTRACT (Maximum 200 Words)

This report presents the results of a study conducted by Genesis Partners LP as part of the United States 

Department of Energy Wind Energy Research Program to develop wind technology that will enable wind systems to 

compete in regions having low wind speeds. The purpose of the program is to reduce the cost of electricity from large 

wind systems in areas having Class 4 winds to 3 cents per kWh for onshore systems or 5 cents per kWh for offshore 

systems. This work builds upon previous activities under the WindPACT project, the Next Generation Turbine project, 

and Phase I of the Low Wind Speed Turbine (LWST) project. This project is concerned with the development of more 

cost-effective gearing for speed increasers for wind turbines.

15. SUBJECT TERMS

gear teeth; low wind speed turbine; 108-kW wind turbine gearboxes; low wind speed; cost-effective gearing for speed 

increasers for wind turbines

16. SECURITY CLASSIFICATION OF: 17. LIMITATION 

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UL

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