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

Efficiency & Renewable Energy, operated by the Alliance for Sustainable Energy, LLC.

Contract No. DE-AC36-08GO28308

Medium-Speed Drivetrain

Test Report

September 1, 2002 — December 30, 2007 

C. Walford, K. Lybarger, T. Lettenmaier, and 

D. Roberts

DNV Renewables (USA) Inc.

Seattle, Washington

NREL Technical Monitor: H. Link

Subcontract Report

NREL/SR-5000-51175

September 2012
NREL is a national laboratory of the U.S. Department of Energy, Office of Energy 

Efficiency & Renewable Energy, operated by the Alliance for Sustainable Energy, LLC.

National Renewable Energy Laboratory

15013 Denver West Parkway

Golden, Colorado 80401

303-275-3000 • www.nrel.gov

Contract No. DE-AC36-08GO28308

Medium-Speed Drivetrain

Test Report 

September 1, 2002 — December 30, 2007

C. Walford, K. Lybarger, T. Lettenmaier, and 

D. Roberts

DNV Renewables (USA) Inc.

Seattle, Washington

NREL Technical Monitor: H. Link

Prepared under Subcontract No. YAM-1-30209-01

Subcontract Report

NREL/SR-5000-51175

September 2012
This publication received minimal editorial review at NREL.

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i

List of Acronyms

AC alternating current

CARB compact aligning roller bearing, registered 

trademark of SKF

CCW counterclockwise

CW clockwise

DAS data-acquisition system

DC direct current

DNG Danish Nacelle Group

DNV Det Norske Veritas

DOE United States Department of Energy

DRS DRS Electric Power Technologies

DTF Dynamometer Test Facility

DTSC drivetrain supervisory controller

EMF electromotive force

FastDAS Fast Data Acquisition System

GRT generator rotor torque

ICP integrated circuit piezoelectric

IGBT insulated gate bipolar transistor

kip kilopound

kN kilonewton

kNm kilonewton meter

kVA kilovolt ampere

NREL National Renewable Energy Laboratory

NWTC National Wind Technology Center

PBL planet bearing load

PSD power spectral density

RTD resistance temperature detector

SCADA supervisory control and data acquisition

SGA sun gear axial displacement

TDC top dead center

THD total harmonic distortion

TTA torque tube axial displacement

TTA-SGD torque tube and sun gear relative axial displacement

TTR torque tube radial displacement

VAC voltage in alternating current
ii

Executive Summary

As part of the National Renewable Energy Laboratory (NREL) WindPACT Advanced Wind 

Turbine Drivetrain Design Study, Det Norske Veritas (DNV) tested a proof-of-concept drivetrain 

prototype using the National Wind Technology Center (NWTC) 2.5-MW Dynamometer Test 

Facility (DTF). The proof-of-concept test article is a full-scale, 1.5-MW integrated drivetrain 

consisting of a single-stage, epicyclic gearbox and close-coupled medium-speed permanent

magnet generator. In addition to the innovative generator design, several developments were 

included in the gearbox design and also were tested. These improvements include a new type of 

bearing for the planet gears, a double-helical ring gear, and an axially floating output pinion.

The test objectives were to characterize the generator and gearbox, and to evaluate thermal 

performance, efficiency, vibration of the drivetrain, and loading in the gearbox. Characterization 

of the generator included measuring the magnitude and frequency content of the output voltage 

and current, and measuring temperatures at various locations on the stator and rotor. The gearbox 

characterization involved measuring the planet gear loads and alignment, the planet bearing 

roller loads, and the sun gear and torque tube displacements. The thermal performance was 

evaluated by measuring the stator temperatures, and measuring the generator coolant and 

gearbox lubrication temperatures and flow rates. Efficiency calculations were performed using 

the input-output method and the calorimetric method. The input-output method involved 

measuring the input mechanical power to the main shaft and the electrical output of the 

generator. The calorimetric efficiency was calculated by summing all of the sources of loss 

coming from the generator in the form of heat. The vibration of the drivetrain was measured in 

three orthogonal directions at two locations by using accelerometers.

During the initial testing of the 1.5-MW integrated drivetrain, one of the stator windings shortcircuited. This required halting the testing and then redesigning the stator windings. The original 

windings were arranged in a parallel configuration and the redesigned windings were arranged in 

an edgewound series configuration.

The drivetrain was operated open-circuit and at several power levels ranging from 400 kW to 

1,500 kW. Most test runs were conducted until the drivetrain achieved a quasi–steady state 

condition and the speed and power were relatively constant. A subset of the test runs was 

conducted until the drivetrain reached thermal equilibrium. The maximum power for which the 

generator was able to achieve thermal equilibrium at less than the design temperature limits was 

1.0 MW.

During the generator characterization testing, the measured current values were greater than the 

design current values, and the measured voltage levels were less than the design voltage values. 

One of the causes of this discrepancy was that the design calculations assumed that the generator 

would operate at unity power factor. The measured power factor of the generator was lagging, 

however, which resulted in an increase in the current. Additionally, the design calculations did 

not account for the reduction in voltage caused by increased magnet temperature. The elevated 

current that resulted from the lagging power factor and reduced voltage caused increased 

conduction losses. This increased magnet temperature and caused further voltage reductions. The 

elevated current also caused additional voltage drop across the stator circuit, which reduced the 

terminal voltage even more. The relationship between voltage, current, and magnet temperature 
iii

forms a positive-feedback system in which even small perturbations can have significant effects. 

The analyses documented in this report were able to account for some—but not all—of the 

discrepancy between the measured and design current and voltage.

In general, the measured generator temperatures were greater than the design values. Comparison 

of the series and parallel winding configurations indicates that the series windings generated

more heat and were less effective at dissipating the loss to the coolant system. At rated power the 

generator windings overheated relatively quickly. Calorimetric efficiency calculations showed 

greater-than-expected losses in the drivetrain. This suggests that the overheating of the generator 

is caused, at least in part, by increased loss. A portion of the increased loss can be attributed to 

the greater-than-expected current levels.

In addition to the calorimetric efficiency calculations, the efficiency of the drivetrain was 

calculated using the input-output method and calorimetric method. Comparison of input-output 

efficiency results from several days of testing demonstrated that input and output power 

measurements are not reliable enough to use to perform efficiency calculations.

The gearbox performed as expected. Ring gear strain measurements showed that the load on the 

ring gear teeth becomes more uniform across a given tooth as the power level increases. 

Measurement of the sun gear and torque tube axial displacement indicated that the spline joint 

between the sun gear and torque tube was effectively isolating the two parts by sliding axially. 

Planet bearing strain measurements showed a relatively linear relationship between mechanical 

load in the bearing and the input power of the gearbox. It also demonstrated that the bearing load 

spreads out circumferentially along the bearing as the load increases.
iv

Table of Contents

List of Figures .............................................................................................................................v

List of Tables.............................................................................................................................vii

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

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

1.2 Scope....................................................................................................................................2

1.3 Objectives ............................................................................................................................2

1.3.1 Component Characterization.......................................................................................2

1.3.2 Thermal Performance..................................................................................................5

1.3.3 Efficiency Tests...........................................................................................................6

1.3.4 Vibration and Acoustic Noise Levels.........................................................................7

2 Test Article Installation.........................................................................................................8

2.1 Test Article Arrangement ....................................................................................................8

2.2 Test Articles.........................................................................................................................8

2.3 Test Facility .........................................................................................................................9

2.4 Test Article Drawings........................................................................................................10

3 Test Instrumentation and Data Acquisition......................................................................10

3.1 Instrumentation ..................................................................................................................10

3.1.1 Electrical System.......................................................................................................10

3.1.2 Motion and Mechanical Loads..................................................................................10

3.1.3 Thermal .....................................................................................................................11

3.1.4 Vibration ..................................................................................................................13

3.2 Data-Acquisition Systems..................................................................................................13

4 Test Campaign ....................................................................................................................14

5 Test Results ........................................................................................................................16

5.1 Component Characterization ............................................................................................16

5.1.1 Generator Characterization ......................................................................................17

5.1.2 Harmonics.................................................................................................................28

5.1.3 Gearbox Characterization .........................................................................................35

5.2 Thermal Performance.........................................................................................................65

5.2.1 Gearbox Fluid Cooling System.................................................................................65

5.2.2 Generator Stator Fluid Cooling System....................................................................65

5.2.3 Generator Rotor Air / Water Cooling System...........................................................74

5.3 Efficiency...........................................................................................................................74

5.3.1 Input-Output Method ................................................................................................75

5.3.2 Calorimetric Method.................................................................................................77

5.4 Vibration Levels.................................................................................................................79

6 Conclusions.........................................................................................................................86

Appendix A. Signal Lists..........................................................................................................87
v 

List of Figures

Figure 1. Cutaway of test article ..................................................................................................... 1

Figure 2. Generator equivalent per-phase circuit diagram .............................................................. 3

Figure 3. Generator phasor diagram for passive rectification, α = 0° ............................................. 4

Figure 4. Drivetrain test setup ......................................................................................................... 8

Figure 5. Gearbox (left); generator assembly power electronic converter (right) .......................... 9

Figure 6. Resistive thermal device locations ................................................................................ 12

Figure 7. Top view of drivetrain with accelerometer coordinates system and accelerometer 

locations labeled .................................................................................................................... 13

Figure 8. Generator open-circuit voltage ...................................................................................... 17

Figure 9. DC-link open-circuit voltage ......................................................................................... 18

Figure 10. Generator current versus rotor torque .......................................................................... 19

Figure 11. Generator voltage versus main shaft speed ................................................................. 19

Figure 12. DC-link voltage versus main shaft speed .................................................................... 20

Figure 13. Generator current versus generator power .................................................................. 21

Figure 14. Generator voltage versus generator power .................................................................. 21

Figure 15. Generator power factor ................................................................................................ 22

Figure 16. Magnet temperature and generator open-circuit voltage ............................................. 23

Figure 17. Stator temperatures and open-circuit generator voltage .............................................. 23

Figure 18. Generator voltage drop with temperature .................................................................... 24

Figure 19. Generator voltage versus generator power .................................................................. 25

Figure 20. Generator current versus generator power .................................................................. 26

Figure 21. Generator current waveforms at one third rated power ............................................... 27

Figure 22. Generator current waveforms at rated power .............................................................. 27

Figure 23. Generator voltage waveforms at one third rated power .............................................. 28

Figure 24. Generator voltage waveforms at rated power .............................................................. 28

Figure 25. Generator current power spectral density at one third rated power ............................. 29

Figure 26. Generator current power spectral density at rated power ............................................ 29

Figure 27. Generator voltage power spectral density at one third rated power ............................ 30

Figure 28. Generator voltage power spectral density at rated power ........................................... 30

Figure 29. Total harmonic distortion of generator voltage and current ........................................ 31

Figure 30. Cogging torque measurement test setup ...................................................................... 31

Figure 31. Rotation scale with tics at 2-mm intervals .................................................................. 32

Figure 32. Static cogging torque ................................................................................................... 33

Figure 33. Generator no-load torque ............................................................................................. 33

Figure 34. Generator current imbalance ....................................................................................... 34

Figure 35. Inverter power factor ................................................................................................... 35

Figure 36. Ring gear strain gage configuration ............................................................................ 36

Figure 37. Ring gear strain gages at X position ............................................................................ 36

Figure 38. Ring gear strain gages at Y position ............................................................................ 37

Figure 39. Ring gear strain gages at Z position ............................................................................ 37

Figure 40. Measuring change in strain level ................................................................................. 38

Figure 41. Tooth load distribution at 1,448 kW ............................................................................ 38

Figure 42. Tooth load distribution from RGZA ............................................................................ 40

Figure 43. Tooth load distribution from RGZB ............................................................................ 40
vi

Figure 44. Tooth load distribution from RGYA ........................................................................... 41

Figure 45. Tooth load distribution from RGYB ........................................................................... 42

Figure 46. Tooth load distribution from RGXA ........................................................................... 42

Figure 47. Tooth load distribution from RGXB ........................................................................... 43

Figure 48. Tooth loading caused by pinion offset ........................................................................ 43

Figure 49. Test article cutaway..................................................................................................... 44

Figure 50. Sun gear and torque tube axial displacement .............................................................. 45

Figure 51. Relative sun gear and torque tube axial displacement................................................. 46

Figure 52. Torque tube radial displacements................................................................................ 47

Figure 53. Sun gear axial position ................................................................................................ 47

Figure 54. Torque tube axial position........................................................................................... 48

Figure 55. Torque tube radial position.......................................................................................... 48

Figure 56. Relative axial displacement between sun gear and torque tube .................................. 48

Figure 57. Sun gear and torque tube axial displacements............................................................. 49

Figure 58. Planetary gear instrumentation location ...................................................................... 50

Figure 59. Planetary gear installed instrumentation ..................................................................... 50

Figure 60. Planet gear axial displacement, lagging ...................................................................... 52

Figure 61. Planet gear axial displacement, outboard .................................................................... 52

Figure 62. Planet gear axial displacement, leading....................................................................... 53

Figure 63. Planet gear axial displacement, lagging ...................................................................... 53

Figure 64. Planet bearing, side and section view.......................................................................... 54

Figure 65. Instrumented planet bearing ........................................................................................ 54

Figure 66. Planet bearing inner race, side view............................................................................ 55

Figure 67. Bearing load calibration fixture................................................................................... 56

Figure 68. CARB bearing calibration, PBL3 (Slot A).................................................................. 57

Figure 69. Planet bearing load (PBL2) at 1,100 kW..................................................................... 57

Figure 70. Planet bearing load (PBL2) at 1,100 kW..................................................................... 58

Figure 71. Planet bearing strain gage peak loads at slot A........................................................... 59

Figure 72. Planet bearing strain gage peak loads at slot B ........................................................... 59

Figure 73. Planet bearing strain gage peak loads, ratio of slot B to slot A................................... 60

Figure 74. Roller load distribution................................................................................................ 61

Figure 75. Planet bearing load distribution................................................................................... 61

Figure 76. Magic Roller waveform (generator power = 1,457 kW)............................................. 62

Figure 77. Magic Roller mean load (generator power = 1,457 kW)............................................. 63

Figure 78. Magic Roller load zone (generator power = 1,457 kW).............................................. 63

Figure 79. Torque response run #1 (rotor speed = 15.5 rpm)....................................................... 64

Figure 80. Torque response run #2 (rotor speed = 15.5 rpm)....................................................... 64

Figure 81. Gearbox lubrication system temperature..................................................................... 65

Figure 82. Winding temperature comparison at 600 kW.............................................................. 66

Figure 83. Winding temperature comparison at 700 kW.............................................................. 67

Figure 84. Temperature run at 1.0 MW........................................................................................ 68

Figure 85. Temperature run at 1.2 MW........................................................................................ 69

Figure 86. Temperature run at 1.5 MW........................................................................................ 70

Figure 87. Design stator temperature breakdown......................................................................... 71

Figure 88. Stator temperature breakdown at 1.0 MW .................................................................. 71

Figure 89. Stator temperature comparison.................................................................................... 72
vii

Figure 90. Stator temperature drops at 1.0 MW ........................................................................... 73

Figure 91. Stator temperature drop comparison ........................................................................... 73

Figure 92. Magnet temperature..................................................................................................... 74

Figure 93. Combined gearbox and generator input-output efficiency by day .............................. 75

Figure 94. Combined gearbox and generator input-output efficiency.......................................... 76

Figure 95. Power electronics and transformer input-output efficiency ........................................ 76

Figure 96. Thermal image example .............................................................................................. 77

Figure 97. Gearbox calorimetric efficiency.................................................................................. 78

Figure 98. Generator calorimetric efficiency................................................................................ 78

Figure 99. Combined gearbox and generator calorimetric efficiency .......................................... 79

Figure 100. Velocity at one third rated power, east...................................................................... 80

Figure 101. Vibration at one third rated power, west ................................................................... 81

Figure 102. Vibration at rated power, east.................................................................................... 82

Figure 103. Vibration at rated power, west................................................................................... 83

Figure 104. Magnitude of gearbox meshing frequency fundamental........................................... 84

Figure 105. Magnitude of gearbox meshing frequency second harmonic.................................... 85

List of Tables

Table 1. Generator Characterization Test Objectives..................................................................... 4

Table 2. Gearbox Characterization Test Objectives....................................................................... 5

Table 3. Thermal Performance Test Objectives ............................................................................. 6

Table 4. Efficiency Test Objectives................................................................................................ 6

Table 5. Vibration Test Objectives................................................................................................. 8

Table 6. Test Article Component Manufacturer and References ................................................... 9

Table 7. Component Dimensions and Weights............................................................................... 9

Table 8. Test System Drawing List............................................................................................... 10

Table 9. Test Type Designations .................................................................................................. 15

Table 10. Test Matrix.................................................................................................................... 16

Table A-1. Instrumentation Connected to the Drivetrain Supervisory Controller and Ported to

the Dyno Data-Acquisition System (1 Hz)........................................................................... 87

Table A-2. Instrumentation Connected to the FastDAS System (2 kHz)..................................... 88

Table A-3. Accelerometer Signals................................................................................................ 90

Table A-4. Calculated Channels................................................................................................... 90
viii
1 

1 Introduction

1.1 Background

The Global Energy Concepts, LLC, later to become DNV Renewables (USA), participated in 

the National Renewable Energy Laboratory (NREL) WindPACT Advanced Wind Turbine 

Drivetrain Design Study (subcontract # YAM-1-30209-01). This participation was divided into 

three phases. Phase I was an in-depth and comprehensive investigation of innovative drivetrain 

designs that potentially could reduce the cost of energy for utility-scale wind turbines. At the end 

of Phase I, a design was selected from the viable candidates. Phase II involved detailed design 

and fabrication of a proof-of-concept version of the selected candidate. Phase III of the study 

involved testing of the final proof-of-concept drivetrain (the test article) using the National Wind 

Technology Center (NWTC) 2.5-MW Dynamometer Test Facility (DTF).

The test article was a full-scale, 1.5-MW integrated drivetrain. Figure 1 shows a threedimensional cutaway of the test article drivetrain, comprised of the gearbox-generator assembly. 

The drivetrain was a single-stage epicyclical gearbox section close-coupled to the medium-speed 

(190 rpm) permanent magnet generator. The test article included a power electronic converter, 

which consists of a purpose-designed passive-diode rectifier and insulated gate bipolar transistor

(IGBT) inverter. The power converter regulates generator torque and converts the variable 

frequency output of the generator power to 60 Hz. A transformer was included in the test setup 

only to match the 480 V output of the power converter with the 575 V service used at the DTF.

Figure 1. Cutaway of test article
2

During the initial testing of the 1.5-MW integrated drivetrain, one of the stator windings shortcircuited. This required halting all testing and redesigning the stator windings, as described in 

1.5 MW Drive Train Generator Coil Failure Event (Global Energy Concepts, LLC, 2006). The 

original windings were arranged in a parallel configuration consisting of many loops of smallgage wire. The redesigned windings were arranged in an edgewound series configuration. The 

edgewound series configuration had much less coil turn-to-turn voltage stress, which improved

robustness. The proof-of-concept article—with the redesigned stator windings—was tested at the 

DTF during the summer and fall of 2007. All results presented in this report are for the 

edgewound series generator configuration, unless otherwise noted.

1.2 Scope

This report covers the testing of the drivetrain system at the DTF. It provides an overview of the 

components that together comprise the entire drive system, and refers to drawing documents that 

define the various mechanical and electrical interfaces. This report briefly describes the test 

equipment, instrumentation, and procedures. A more detailed description of the test setup and 

procedure is provided in document N60006, WindPACT Drive Train Test Plan. The 

interconnection and integration of the various system components also is covered in a separate 

document, N60016, WindPACT Drive Train Commissioning Plan (Figure 1).1

1.3 Objectives

Within the overall goal of demonstrating and proving the capabilities of the DNV WindPACT 

drivetrain, several specific objectives motivated the test campaign, including:

• Component characterization,

• Thermal performance,

• Efficiency, and

• Vibration and acoustic noise level.

1.3.1 Component Characterization

The drivetrain is a proof-of-concept first article; therefore, assumptions about the operating 

characteristics of the gearbox, generator, and power electronic units were made during the design 

phase. The test campaign refined these assumptions by measuring internal characteristics such as 

deflection, impedance, and temperature rise as functions of load, speed, and time.

1.3.1.1 Generator Characterization

The WindPACT drivetrain includes a purpose-designed synchronous generator. The generator 

has a radial-flux permanent magnet design. The stator consists of 72 individual (salient) poles 

connected in four discrete quadrants; each quadrant includes all three phases in a “star”

arrangement with an un-grounded neutral. The stator coils are surrounded by a water jacket to 

provide cooling. The rotor consists of 48 permanent-magnet poles mounted to the outer rim of a 

flanged disc. The rotor is mounted inside the stator and its nominal maximum rotational speed 

during normal operation is 191 rpm. The equivalent per-phase circuit diagram of the permanent

magnet synchronous generator is shown in Figure 2. In this diagram, the assumed current flow 

follows the standard convention for a motor. The actual current flow for a generator will be 

negative if using this convention.

 1 Documents N60006 and N60016 are unpublished internal NREL reports. 
3

• Emq—Internal electromotive force or open-circuit voltage

• Xs—Synchronous reactance

• Rs—Stator resistance

• Rc—Equivalent core loss

• Vst—Terminal voltage

Figure 2. Generator equivalent per-phase circuit diagram

Near the end of Phase II, a significant design change with regard to the power electronics was 

initiated. This change was required due to the adoption of grid codes (in particular, the lowvoltage ride-through requirement) that the drivetrain would have had difficulty meeting if using

the power electronics topology originally selected in Phase I. As a result of the new requirement, 

the NREL and DNV project team decided to replace the thyristor-based line-commutated current 

source converter selected in Phase I with an IGBT-based voltage-source converter. The converter 

utilized by Clipper Windpower in testing the DGD1 drivetrain at the DTF in 2003 was modified 

in October 2005 to enable its integration with the DNV drivetrain.

The output of the generator is connected to four diode-bridge rectifiers that provide DC current 

to four IGBT-based voltage-source inverters operated in parallel. Each 375-kW inverter 

accommodates one quadrant of the generator. The performance of the generator when connected 

to this passive rectifying power electronics system largely is determined by the generator internal 

electromotive force (EMF) and synchronous reactance. The relationship between these 

parameters and the per-phase voltages and currents is determined using the phasor diagram 

shown in Figure 2.

+

Emq

-

Xs

Ec

I

s

Vst

+

-

I

st

I

c Rc

Rs
4

d

q

I

sd

I

s

I

st

I

sq

I

c

Vst

Emq

I

st*Rs

j Xs*I

s

j Xs*I

sq

j Xs*I

sd

Ec

Ø

Figure 3. Generator phasor diagram for passive rectification, α = 0°

The parameters used in the circuit model shown in Figure 3 were measured during the 

acceptance tests performed at the assembly factory. The acceptance tests included measuring the 

“static” cogging torque, measuring the open-circuit voltage, validating the stator insulation 

scheme, and initially verifying the thermal performance of the stator cooling system under DC 

load. The generator was not run dynamically under load during these tests. The test campaign

carried out at the DTF measured the performance of the generator when coupled to the 

rectifier/inverter system. Table 1 lists the objectives of the test campaign relative to the generator 

and power electronic system.

Table 1. Generator Characterization Test Objectives

Objective Measurement

Measure generator electrical performance AC voltage and current at generator terminals and DC voltage at 

rectifier terminals as functions of rotor torque and speed

Quantify generator electrical characteristics Voltage and current waveforms at discrete operating points

Measure cogging torque pulsations Generator torque variation at low speed

Characterize generator torque command 

transient response Generator current and voltage waveforms, rotor torque time series

Quantify current balance between all 4

generator quadrants Quadrant currents at discrete operating points

Evaluate converter output quality Converter output power factor
5

1.3.1.2 Gearbox Characterization

The WindPACT drivetrain gearbox is a single-stage planetary (i.e., epicyclic) design with 

double-helical gearing. Helical gearing is preferred over spur gearing in gearboxes because it 

provides increased contact ratio, thereby increasing the load-carrying capacity and reducing 

noise generated by tooth meshing. Double-helical gearing—as opposed to single-helical—

supports the sun and planet gears with the mesh itself, rather than relying on additional thrust 

bearings to react the axial load induced by the helix angle.

The challenge with the double-helical arrangement is that the gearing accuracy is critical to 

correct operation. Although tooth-to-tooth, profile, and lead accuracy are important for proper 

load-sharing in any planetary gearbox design, the double-helical arrangement is more 

demanding. Any deviation of the apexes (the theoretical intersection of each left and right helix) 

affects the ability of the planets to share the torque load equally.

The gearbox was manufactured with scrupulous attention paid to apex run-out. Additionally, a 

unique bearing was used to mount the planets to the carrier pins. This “toroidal” bearing enables 

the planet gear to float axially, thus allowing each planet to “follow” any run-out in the ring gear 

apex. This bearing also allows the planet to tilt with respect to the carrier pin so that it can 

accommodate any misalignment—due to errors or deflection—between the planet carrier and the 

ring gear.

The sun gear in this planetary arrangement is floating; it is held in position only by the planets 

meshing. Any misalignment motion of the planet gears also affects the position of the sun gear 

axially and radially. To accommodate this motion, the sun gear is connected to the generator by 

means of a torque tube that is compliant in the axial and radial (but not torsional) directions. The 

connection between the sun gear and torque tube is a sliding spline joint, which also allows each 

to move axially relative to the other. One of the goals of the test campaign was to characterize 

the operation of the planetary section of the drivetrain. Results from these characterizations 

enable evaluation of the effectiveness of the design techniques. The objectives are defined in 

Table 2.

Table 2. Gearbox Characterization Test Objectives

Objective Measurement

Evaluate torque load share among planet gears Planet bearing loads, ring gear tooth strain, and generator 

torque as functions of rotor torque and speed

Evaluate dynamic planet gear alignment with ring 

gear apex

Ring gear strain on each helix, displacement of planet and 

sun gears as functions of rotor torque and speed

Evaluate roller load distribution in planet bearing Strain measurements in planet bearing as functions of rotor 

torque and speed

Evaluate torque tube/sun gear spline joint motion Sun gear and torque tube displacements

1.3.2 Thermal Performance

The drivetrain includes a circulating lubrication system for the mechanical components with an 

integrated air cooler to remove heat from the oil. A variable flow-rate fluid system provides 

circulating coolant for the generator stator and rotor. One of the objectives of the test campaign 
6

was to verify that these systems operate as intended, and maintain the system temperatures 

within the operational limits. Thermal performance test objectives are included in Table 3.

Table 3. Thermal Performance Test Objectives

Objective Measurement

Evaluate gearbox fluid cooling system Planet bearing, main bearing race, main bearing fluid, 

gearbox fluid inlet and return temperatures

Evaluate generator stator fluid cooling system Selected stator coil temperatures, coolant inlet and outlet 

temperatures

Evaluate generator rotor air/water cooling system Rotor magnet temperatures

1.3.3 Efficiency Tests

The overall goal of the study was to identify and demonstrate a drivetrain with improved cost of

energy. An important objective therefore was quantifying the efficiency of the individual 

components and of the drivetrain system as a whole. Efficiency was measured at several points 

along the design torque-speed curve. Efficiency test objectives are included in Table 4.

Table 4. Efficiency Test Objectives

Objective Measurement

Quantify mechanical efficiency Coolant flow rates and temperature differentials; torque and speeds

Quantify generator efficiency Generator torque and speed; output current and voltage

Efficiency is estimated using two distinct methods, the input-output method and the calorimetric 

method, as described below.

1.3.3.1 Input-Output Method

This method is defined in IEEE 115-1995, section 4.6.2, for motors and generators and also is 

applicable to the gearbox and power electronics unit. The subsystem input power and the output 

power are measured, then efficiency is calculated as the ratio of the output power to the input 

power.

Efficiency (%) = 100 ∙ (output power) / (input power)

For the gearbox, both the input and output power are calculated from the product of the main 

shaft and generator rotor shaft torques and speeds. For the generator, input power is calculated 

from the product of the generator rotor shaft torque and speed. The output power is calculated 

from the sum of the products of the four quadrant voltages and currents.

1.3.3.2 Calorimetric Method

This method is described in IEEE 115-1995, section 4.5 (available at 

http://standards.ieee.org/findstds/standard/115-1995.html). The output power and the 

calorimetric (heat) losses are measured. Power is calculated in same manner as for the inputoutput method. 
7

Efficiency is calculated using the equation below.

Efficiency (%) = 100 – (losses ∙ 100) / (output power + losses)

The losses are defined as follows.

Loss (kW) = 0.264 ∙ cpc ∙ Gc ∙ Qc ∙ (thc - tcc)

• cpc—The specific heat of the coolant (relative to water) at average temperature thc and tcc

• Gc—The specific gravity of the coolant (relative to water) at average temperature thc 

and tcc

• thc —The outlet coolant temperature ( ̊C)

• tcc—The inlet coolant temperature ( ̊C)

• Qc—The rate of coolant flow in gallons per minute

Both the generator and gearbox have distinct fluid systems, but these components essentially are 

housed in the same steel structure. Undefined thermal short-circuits in the housing structure, 

however, make it impossible to assign all of the heat energy absorbed by a particular fluid 

system to the losses associated with that device. The error associated with this “cross-talk” is 

difficult to estimate. The efficiency calculation for the generator/gearbox assembly is more 

accurate than the calculations used for each component. The calorimetric tests require that the 

heat transfer rates reach steady state. This was achieved by operating the test articles until the 

rates of change of the fluid temperatures and the structure or heat-sink temperature fall below an 

established value.

The drivetrain installation includes specific instrumentation at several locations to support 

efficiency measurements, including:

• Main shaft torque and speed;

• Generator torque and speed;

• Generator output current and voltage;

• Inverter output current and voltage;

• Lube system flow rate and temperatures; and

• Generator coolant flow rate and temperatures.

1.3.4 Vibration and Acoustic Noise Levels

The test plan calls for vibration and acoustic emission measurements of the gearbox/generator 

assembly to be made at selected points on the design torque-speed curve. The ambient noise of 

the DTF during testing (e.g., dynamometer, ventilation fans, lubrication pump), however, makes 

performing accurate sound-pressure levels difficult. Although some experimentation was 

performed with acoustic noise–level measurements, the acoustic emissions of the drivetrain were 

not evaluated. Vibration measurements are described in ANSI/AGMA 6000-B96 (available from 

ANSI at http://webstore.ansi.org/RecordDetail.aspx?sku=ANSI/AGMA%206000-

B96%20(R2002)). Table 5 lists the vibration test objectives. Instrumentation for measuring 

vibration includes accelerometers mounted on the gearbox to measure longitudinal, lateral, and 

vertical acceleration.
8

Table 5. Vibration Test Objectives

Objective Measurement

Evaluate drivetrain vibration levels Linear acceleration

2 Test Article Installation

2.1 Test Article Arrangement

The test setup is shown in Figure 4. The drivetrain assembly, consisting of a single-stage gearbox 

integrated with a medium-speed permanent magnet generator, is driven by the dynamometer 

drive system. Each of the four three-phase windings of the generator is connected directly to a 

passive rectifier (three-phase diode bridge). Each rectifier output feeds one IGBT-based inverter 

bridge. The direct current (DC) link operates at a variable voltage which is dependent on 

generator speed. Generator current—and hence torque—is controlled by setting an alternating 

current (AC) output current to each inverter bridge. The inverters have been designed to operate 

at unity power factor only. Two 900 kVA output transformers provide voltage-matching between 

the 480 VAC inverter output and the 575 VAC utility source at the DTF.

900 kVA

480V Y: 575V ZZ

900 kVA

480V Y: 575V ZZ

To NREL 

575V Panel

To NREL 

575V Panel

Transformer #2

Transformer #1

Inverter Enclosure #2

Rectifier Enclosure Inverter Enclosure #1

Single Stage 

Gearbox (1:9.3)

Permanent 

Magnet 

Synchronous 

Generator

To NREL 

Dynamometer Drive

Figure 4. Drivetrain test setup

2.2 Test Articles

The articles tested are listed below and shown in Figure 5. Only one of each article was tested.

• Drivetrain assembly, consisting of a single-stage planetary gearbox (ratio 9.3: 1) and a 

1,500-kW, 191-rpm, 690-volt permanent magnet generator, assembled on a test stand.

• Power electronic converter, consisting of a single rectifier cabinet containing four threephase diode bridges, and two identical inverter cabinets, each containing two 

independently controlled voltage-source inverter bridges.
9

The manufacturer and reference specification or drawing number for each component is listed in 

Table 6. Dimensions and weights for all components are listed in Table 7.

Table 6. Test Article Component Manufacturer and References

Component Manufacturer Specification or Reference

Drivetrain assembly Brad Foote Gear Works;

DRS Electric Power Technologies

Mechanical: Powertrain Engineers, Inc. (PEI) dwg. 0202-

10100-80; Generator: DNV specification N60004 Rev B

Test stand Lindquist Manufacturing Corp. Powertrain Engineers, Inc. dwg. DNV-0202-10205 Rev 6

Power electronic 

converter Xantrex Technology, Inc. Xantrex Schematic Drawing #151858

900-kVA 

transformers Mag-Tran Equipment Corp. DNV specification N60009

Table 7. Component Dimensions and Weights

Component Length m (ft) - Width m (ft) - Height m (ft) Weight kg (lb)

Gearbox/generator assembly 2.5 (8.2) 2.0 (6.5) 2.0 (6.5) 24,600 (54,120) *

Test stand 2.0 (6.6) 3.5 (11.5) 1.0 (2.2) 5,800 (12,760)

Rectifier cabinet 1.0 (3.1) 0.7 (2.3) 2.5 (7.7) 410 (900)

Inverter cabinet (each) 1.4 (4.3) 0.7 (2.3) 2.5 (7.7) 1,500 (3,300)

900-kVA transformers 2.4 (8) 2.1 (7) 3.0 (10) 5,910 (13,000)

* Note that gearbox section weighs 13,100 KG (28,820 lb); generator section weighs 11,500 kg (25,300 lb) ().

Figure 5. Gearbox (left); generator assembly power electronic converter (right) (Photo credit: 

Chris Walford, DNV Renewables Inc.)

2.3 Test Facility

The tests were performed at the NREL 2.5-MW Dynamometer Test Facility located at the 

NWTC in Boulder, Colorado. The DTF is capable of driving the test article either at a defined 

fixed rotational speed, or at a nominal rotational speed with a superposed sinusoidal variation. 

The DTF includes a digital data-acquisition system, the “Dyno DAS.” This system was used for 

short-term measurements of mechanical and electrical parameters at sample rates in the 1-kHz to 
10

2-kHz range. Measurement of slowly varying parameters such as temperature, flow, and pressure 

were recorded by the DNV drivetrain supervisory controller (DTSC).

2.4 Test Article Drawings

The general arrangement, installation, and interconnections of the test articles are defined in the 

drawings listed in Table 8.

Table 8. Test System Drawing List

Drawing # Title Owner

B-N6-10054 General Arrangement DNV

B-N6-10051 Sensor and Control Installation DNV

B-N6-10052 Electrical Schematic DNV

0202-10227-80 DNV Load Test Layout PEI

0202-10100-80 Medium-Speed Turbine Assembly PEI

No number 

assigned Test System Wiring Diagram BEW

3 Test Instrumentation and Data Acquisition

3.1 Instrumentation

To meet the objectives of the test plan, instrumentation was installed on the drivetrain to measure 

a variety of parameters including voltage, current, displacements, mechanical load, temperature, 

and vibration.

3.1.1 Electrical System

The generator, rectifiers, and inverters were instrumented with current and voltage transducers to 

measure current and voltage levels, power flow, power factor, and frequency content. The 

electrical instrumentation included the following.

• High-bandwidth current transducers on all phases from each generator quadrant

• High-bandwidth line-to-neutral voltage transducers on all phases from each generator 

quadrant

• High-bandwidth voltage transducers on output terminals of the rectifier bridge connected 

to each generator quadrant

• Power transducers on the output terminals of each inverter

• High-resolution position transducer on the generator rotor

3.1.2 Motion and Mechanical Loads

In addition to the torque transducer integrated into the DTF drive line, purpose-built instrumentation was incorporated into the drivetrain to measure the torque and speed of the main shaft and 

generator. Displacements and strain were measured with the following instrumentation.

• Four proximity probes at equal quadrants on one of the carrier pins to measure the axial 

displacement of the planet gear relative to the carrier pin (PGD1 through PGD4).

• Two strain-gauged slots in the inner race of one of the planet bearings to measure the 

strain as the loaded roller passes over them (PBL1 through PBL8).
11

• One proximity probe measuring radial displacement of the torque tube (TTR) and another 

measuring axial displacement of the torque tube (TTA), relative to the gearbox case.

• One proximity probe measuring axial displacement of the sun gear (SGA) relative to the 

gearbox case.

• Three pairs of strain-gage bridges measuring root strain along adjacent left-helix and

right-helix teeth in the ring gear. These are relative measurements and are not calibrated 

to a particular load.

• A torsional strain gage in the torque tube to measure generator rotor torque (GRT) (highspeed torque). This has been laboratory calibrated to the torque load.

3.1.3 Thermal

Instrumentation was installed on the drivetrain to measure the generator winding, drivetrain case, 

and fluid temperatures. Thermal imaging also was used for measuring temperatures during and 

after particular test runs.

3.1.3.1 Generator Winding Temperature

To monitor the winding temperatures and flow of heat from the stator windings to the generator 

coolant, several stator temperatures were measured. This was accomplished through the 

installation of four sets of resistance temperature detectors (RTDs) on and around the windings at 

different locations on the stator. Each set of RTDs consists of five RTDs installed at the 

following positions (see Figure 6).

• Outside of the end-turn

• Inside of the end-turn

• Laminate tooth

• Back iron

• Cold plate
12

Figure 6. Resistive thermal device locations

Surface-mounted RTDs measured the temperatures at the outside of the end-turn, the inside of 

the end-turn, the laminate tooth, and the back iron. The cold plate temperatures were measured 

by embedding RTD probes approximately 1 inch deep into the cold plate.

3.1.3.2 Magnet Temperature

The magnet temperatures were measured directly using infrared thermometry and thermocouples, and indirectly based on changes in the generator open-circuit voltage. Access ports were 

installed in the generator cover for directly measuring the rotor magnet temperatures. This 

required stopping the drivetrain and manually measuring the temperature. There is linear 

relationship between open-circuit voltage and magnet temperature, therefore the magnet 

temperatures were estimated by operating the generator open-circuit immediately before or after 

particular test runs.

3.1.3.3 Drivetrain Case Temperature

The temperature of the drivetrain case was measured using a surface-mount RTD as well as 

thermal imaging. To measure planet-bearing and main-bearing temperatures, sensors also were 

embedded in the gear case. To estimate the thermal heat loss of the drivetrain, a heat-flux sensor 

was installed next to the RTD on the drivetrain case. The heat-flux sensor consists of a thin film 

that functions as a thermopile transducer. To improve the signal-to-noise ratio, the heat-flux 

sensor is positioned on one of the hottest portions of the case. A linear relationship is assumed 

between case temperature and heat flux, which enables the measured heat flux to be scaled with 

temperature to approximate the average heat flux of the drivetrain.

Thermal images of the drivetrain were used to determine the average temperature of the 

generator and gearbox to account for thermal variation in the drivetrain case. The thermal images 

also enable analysis of the heat flow through salient features such as the drivetrain stand.
13

3.1.3.4 Fluid Temperature and Flow

Temperature sensors were installed in the lubrication and coolant lines to measure supply and 

return temperatures. Flow and pressure sensors also were installed in the lubrication and coolant 

systems. The fluid temperature measurements were used to ensure that the drivetrain was 

operating within acceptable limits. These fluid measurements were combined with the flow 

measurements to calculate heat loss.

3.1.4 Vibration

The linear vibration of the drivetrain was measured in three orthogonal directions, X, Y, and Z, 

as shown in Figure 7 (note that Z is coming out of the page). The X direction is defined along the 

axis of the drivetrain shaft, with X increasing from rotor to generator. The Y direction is defined 

as perpendicular to the X direction and in the horizontal plane, with Y increasing to the left of 

the generator when looking into the rotor. The Z direction is defined as perpendicular to the X

and Y directions and in the vertical plane, with Z increasing upward.

Two vibration sensors—each consisting of three accelerometers, and capable of measuring 

vibration in the X, Y, and Z directions—were attached to the exterior of the drivetrain case over 

the bearings. The vibration sensors were fastened to the casing by gluing on small pads that had

threaded holes and then screwing the sensors into the pads.

Figure 7. Top view of drivetrain with accelerometer coordinates system and

accelerometer locations labeled

3.2 Data-Acquisition Systems

Most of the instrumentation is integrated into the test articles. These signals are accessible to the 

Dyno DAS at several locations as defined in document B-N6-10051, Sensor Installation, and 

document B-N6-10052, Electrical Schematic. Other instrumentation is incorporated into the 

DTF, or is portable laboratory equipment.
14

Four measurement systems ran simultaneously during the test:

• Dynamometer Test Supervisory Control;

• Dynamometer Test Facility Supervisory Control and Data Acquisition (SCADA);

• Dynamometer Test Facility Fast Data Acquisition System (FastDAS); and

• Dynamometer Test Facility Vibration Monitoring System (Vibration Monitor).

All four of these systems are based on National Instruments technology and run under the 

LabVIEW platform. The DataSocket interface was used to port signals between the DTSC, 

SCADA, and Fast DAS.

The DTSC ran continuously and recorded temperatures, fluid pressures, and flow rates of all of 

the critical electrical and mechanical components. These signals change relatively slowly, so 

they were sampled at 1 Hz. The signals were converted to engineering units in real time and 

displayed on the DTSC operator screen, but were not recorded by the DTSC. Appendix A, 

Table A-1, lists the sensors connected to the DTSC. The SCADA recorded the signals measured 

by the DTSC. FastDAS recorded the signals that require a high sampling rate. These were 

sampled at 2 kHz for short periods (less than one minute) to limit the size of the data files 

(Appendix A, Table A-2, lists these signals). Refer to drawing B-N6-10052, Electrical 

Schematic, for details on the electrical connections.

The vibration monitor recorded the accelerometer output at 2 kHz. Appendix A, Table A-3, lists 

the accelerometer signals. A vibration monitor calculates velocity and displacement from the 

accelerometer data. Some calculated channels also were required for monitoring drivetrain 

condition and status during the test campaign. These were implemented in “real-time” in the 

SCADA. These channels are defined in Appendix A, Table A-4. All of these signals were

recorded at 2 kHz along with the measured signals. Additional portable or stand-alone instruments were required for some tests, such as for infrared temperature measurements and ongoing 

fluid-cleanliness checks. These are identified in the description of the particular test procedure.

4 Test Campaign

The DNV WindPACT drivetrain is a prototype system that was commissioned for the first time 

when it was installed at the DTF. This commissioning process, described in DNV document 

N60016, Test System Commissioning Plan, included installation and alignment to the 

dynamometer, connecting power and signal wiring, connecting the lubrication and fluid cooling

systems, and power-up of the rectifier/inverter system.

During the testing phase, the test article gradually was brought up to rated speed and power in 

incremental steps. To ensure that the systems were operating as intended, the test article 

performance was evaluated at each step prior to beginning the next step. This approach ensured

that if a problem with the test article developed and testing had to cease for a period, then some 

test data would be available and analysis could proceed in the interim.

This approach proved to be wise when the generator winding failure occurred after only a brief 

period of operating at near rated power. This failure was not entirely unanticipated, because 

similar failures had occurred during the factory test of the original coil design. At that time, the 
15

coils were rewound with a more robust insulation scheme and successfully retested. The risk of 

an additional failure during the actual test was too great, however, and it was decided that the 

coils would be reconfigured. The generator rewind required disassembly of the test article in the 

DTF, shipment of the core assemblies to the rewind facility, and reassembly using the new coils. 

The entire process halted the test campaign for four months. The test campaign then was 

restarted as if the unit was a new test article, because the differences in the generator design were 

significant both electrically and thermally.

Overall, there were five test types; each was given a numeric designation as shown in Table 9. 

The basic test parameters for each test—speed and power—are listed in Table 10. The number in 

each test designation describes the type of test, and the letter corresponds to a specific torquespeed operating point. Except for the last test listed in the matrix, all of the tests were steadystate. The test article was brought to a constant speed and then enabled to generate at the target 

output power level. Note that the power level actually was set by a current command—in 

amps—to the Xantrex inverters. In practice this command was adjusted by trial-and-error to 

achieve the target output power level at constant speed. The service voltage at the DTF 

fluctuated during testing but generally was low. The power converter of the test article is current 

limited, therefore reduced service voltage reduces the allowable output power. The Stage 1 tests 

(levels a through c) shown in Table 10 are shaded grey because it was not possible to reach the 

target output power level due to insufficient bus voltage.

Table 9. Test Type Designations

Type Title

1 Open-circuit DC voltage characterization (cold)

2 Power curve operation

3 Open-circuit DC voltage characterization (hot)

4 Efficiency

5 Vibration and acoustical
16

Table 10. Test Matrix

5 Test Results

5.1 Component Characterization

During all drivetrain testing the dynamometer was operated using speed control. Using this 

function, the desired speed is specified in the dynamometer controller and the torque applied by 

the dynamometer is adjusted to maintain the desired speed. The output of the inverters connected 

to the generator was controlled by specifying the desired output current in the inverter controller. 

For a given current command, the output power of the inverters fluctuated based on changes in 

the grid voltage. The generator power was determined by setting the desired speed in the 

dynamometer and the desired current in the inverters. The current command values for the 

inverter current that correspond with the operating points in the test matrix were determined 

empirically. During the commissioning of the drivetrain, the inverter current set points were 

determined by operating the generator at each speed in the test matrix and adjusting the inverter 

current until the output power matched the value stipulated in the test matrix.

Stage

Main 

Shaft 

Speed 

(rpm)

Generator Shaft 

Speed 

(rpm)

Drivetrain 

Output 

Power 

(kW)

OpenCircuit 

Cold Test#

Power 

Curve 

Test #

OpenCircuit Hot 

Test #

Efficiency 

Test #

Vibration & 

Acoustical 

Test #

Series 1 Series 2 Series 3 Series 4 Series 5

1

12.2 114 233 1a — — — —

12.4 116 301 1b — — — —

13.4 124 378 1c — — — —

2

14.3 133 467 1d 2d - 4d 5d

15.4 142 568 1f 2f 3f — —

3

16.2 151 682 1g 2g — 4g 5g

17.2 161 811 1h 2h — — —

18.1 170 952 1j 2j 3j 4j 5j

4

19.1 178 1,110 1k 2k 3k — —

20.1 187 1,284 1m 2m 3m 4m 5m

20.5 191 1,460 1n 2n 3n 4n 5n

20.5 191 1,500 1p 2p 3p 4p 5p

—

Ramp 0 to 

191 @ 

10 rpm 

per second

0 — — — — 5r
17

5.1.1 Generator Characterization

5.1.1.1 Electromotive Force Voltage Constant

The open-circuit voltage of the generator, with both the original parallel windings and the 

reworked series windings, was slightly greater than expected. Figure 8 is a plot of the generator 

open-circuit voltage. It was not possible to operate the generator open-circuit faster than 169 rpm 

due to protective voltage limits in the power electronics. The parallel windings generated an 

open-circuit voltage that was 3.5% greater than that specified in the design, and the series 

windings generated an open-circuit voltage that was 4.4% greater than specified in the design.

Figure 8. Generator open-circuit voltage

The DC-link voltage is measured at the output of the rectifier. Figure 9 is a plot of the DC-link 

open-circuit voltage of the power electronics as a function of the generator speed. The series and 

parallel winding configurations produced similar open-circuit voltages in the DC-link.

0

100

200

300

400

500

600

700

800

900

130 135 140 145 150 155 160 165 170 175

Generator Speed (RPM)

Generator Open-Circuit 

Voltage (V) 

Parallel Winding

Series Winding

Design
18

Figure 9. DC-link open-circuit voltage

Figure 10 is a plot of the design and measured generator current as a function of rotor torque. All 

references to current in this document refer to the total generator current, which is calculated by 

summing the average current of each quadrant. At low torque levels the measured current is

relatively close to the design values. At greater torque values, however, the measured generator 

current is significantly greater than that specified in the design.

The test matrix in Table 10 defines the operating envelope of the drivetrain. In this test matrix, 

increased power is achieved by simultaneously increasing torque and speed. Consequently, all 

plots of torque also have varying speed, and all plots of speed also have varying torque. Speed 

was not held constant for the measurements in Figure 10, and torque was not held constant for 

the measurements in Figure 11.

0

200

400

600

800

1000

1200

1400

130 135 140 145 150 155 160 165 170 175

Generator Speed (RPM)

DC-Link Open-Circuit 

Voltage (V)

Parallel Winding

Series Winding
19

Figure 10. Generator current versus rotor torque

Figure 11 is a plot of the measured and design generator voltage as a function of main shaft

speed. The measured generator voltage is near the design value at low speeds but drops to less 

than the design value at greater speeds. The discrepancy between the two voltage values 

propagates to the DC-link voltage, as shown in Figure 12. The voltage data in Figure 11 and 

Figure 12 has a large amount of scatter, indicating that the generator voltage is being influenced 

by a factor other than generator speed. 

Figure 11. Generator voltage versus main shaft speed

0

200

400

600

800

1000

1200

1400

1600

200 300 400 500 600 700 800 900

Rotor Torque (kN-m)

Generator Current (A)

Design

Measured

0

100

200

300

400

500

600

700

800

900

12 14 16 18 20 22

Main Shaft Speed (RPM)

Generator Voltage (V)

Design

Measured
20

Figure 12. DC-link voltage versus main shaft speed

Although there are correlations between current and torque and between voltage and speed, the 

plots of current and voltage provided above could be misinterpreted because speed is not held 

constant during period shown in the torque plots and torque is not held constant in the speed 

plots provided. The generator current and voltage are plotted against generator power in 

Figure 13 and Figure 14. Measured results are presented for the original parallel winding 

configuration and for the reworked series winding configuration. The measured generator current 

and voltage for both winding configurations is consistent with the design values at low power 

levels but diverges at greater power levels. There is scatter in the current data—despite the tight 

regulation of the inverters—because variations in voltage affect the power of the generator. The 

scatter in generator voltage data increases dramatically as the power level increases. Some of the 

discrepancy between the measured and design values of current and voltage and the scatter of the 

voltage data can be explained by taking into account the power factor and thermal effects.

0

200

400

600

800

1000

1200

12 14 16 18 20 22

Main Shaft Speed (RPM)

DC-Link Voltage (V)

Design

Measured
21

0

200

400

600

800

1000

1200

1400

1600

400 600 800 1000 1200 1400 1600

Generator Power (kW)

Current (A)

Design

Measured Series

Measured Parallel

Figure 13. Generator current versus generator power

0

100

200

300

400

500

600

700

800

900

400 600 800 1000 1200 1400 1600

Generator Power (kW)

Voltage (V)

Design

Measured Series

Measured Parallel

Figure 14. Generator voltage versus generator power

Figure 15 is a plot of the generator power factor. This factor ranges from 0.93 to 0.96, which is 

consistent with the theoretical power factor of a passive rectifier. The design assumed that the 

generator would operate at unity power factor. Operation at a power factor of 0.93 to 0.96 results 

in 4% to 7% more current than operation at unity power factor. The lagging power factor of the 

generator accounts for some of the discrepancy between the design and measured current in 

Figure 13. The lagging power factor, however, does not explain why the measured generator 
22

current diverges from the design current as the power level increases. There is very little scatter 

in the data points of Figure 15; power factor is not causing the scatter present in the current and 

voltage data in Figure 13 and Figure 14.

Figure 15. Generator power factor

The theoretical temperature coefficient of the generator electromotive force is -0.11% per degree 

Celsius. Therefore for every 1°C increase in magnet temperature the open-circuit generator 

voltage should decrease by 0.11%. Figure 16 provides a plot of the measured and calculated 

magnet temperature as a function of generator open-circuit voltage. The calculated magnet 

temperature is based on the theoretical temperature coefficient of the magnets. To measure the 

magnet temperature, the drivetrain was stopped immediately after an open-circuit test and the 

temperature was measured through an access port using a temperature probe. This measurement 

method has inherent inaccuracy due to the great amount of heat radiating from the windings, 

which are designed to operate at 165°C. The proximity of the magnets to the windings and the 

lack of convection caused the magnets to heat up after the generator was stopped. Using the 

open-circuit voltage to estimate the magnet temperature is more accurate than directly measuring 

the magnet temperature through the access ports. Consequently, the magnet temperatures 

calculated using the open-circuit voltage in Figure 16 better reflect the actual magnet 

temperature than do the temperature-probe results.

0.90

0.91

0.92

0.93

0.94

0.95

0.96

0.97

0.98

0.99

1.00

400 600 800 1000 1200 1400 1600

Generator Power (kW)

Power Factor
23

Figure 16. Magnet temperature and generator open-circuit voltage

Figure 17 is time-series plot of the open-circuit generator voltage and stator temperatures. As the 

stator temperatures increase, the generator voltage decreases. The decrease in generator voltage 

as temperatures increase is prevalent in all open-circuit and loaded test runs.

Figure 17. Stator temperatures and open-circuit generator voltage

30

40

50

60

70

80

90

100

110

120

130

440 450 460 470 480 490

Generator Open-circuit Voltage (V)

Magnet Temperature (deg C) 

Measured

Calculated

0

30

60

90

120

150

6:30 AM7:00 AM7:30 AM8:00 AM8:30 AM9:00 AM9:30 AM

Time

Temperature (deg C)

0

200

400

600

800

1000

Voltage (V)

Average outside end-turn temp.

Average inside end-turn temp.

Average laminate tooth temp.

Average backiron temp.

Average cold plate temp.

Avgerage generator voltage
24

Figure 18 shows the measured relationship between change in generator open-circuit voltage and 

change in stator temperature for the test run illustrated in Figure 17. The temperature change is 

relative to the ambient temperature, and the voltage change is relative to the voltage at the 

ambient temperature. The generator is operating at no load and without the coolant loop running.

It therefore it seems reasonable that the magnet temperatures and stator temperatures would 

change at approximately the same rate. Based on Figure 18, the empirical temperature coefficient 

of the generator voltage is -0.11% per degree Celsius. Assuming that the change in temperature 

of the stator is approximately equal to the change in temperature of the magnets, the theoretical 

temperature coefficient of the generator voltage is confirmed.

y = -0.0011x + 0.0194

R2

 = 0.9988

-0.09

-0.08

-0.07

-0.06

-0.05

-0.04

-0.03

-0.02

-0.01

0.00

0 10 20 30 40 50 60 70 80 90 100

Change In Average Stator Temperature (deg C)

% Change in Generator Voltage

% V

% V Treadline

Figure 18. Generator voltage drop with temperature

For a subset of the data points in Figure 13 and Figure 14, the open-circuit voltage of the 

generator was checked immediately after the operating data was collected. The open-circuit 

voltage then was used to estimate the reduction in generator voltage caused by increased magnet 

temperature. The design voltage and current values did not take into account the effect of 

elevated magnet temperature. Figure 19 and Figure 20 contain the subset of current and voltage 

data points for which the open-circuit voltage was checked immediately after the test run. 

Figure 19 and Figure 20 also contain adjusted versions of measured current and voltage values. 

The measured current and voltage values have been adjusted to compensate for the effects of 

power factor and temperature.

The adjustment of the voltage and current was performed based on the following set of 

assumptions.

• The design calculations assumed the generator would operate at unity power factor. 

During testing, the generator operated at a power factor of 0.93 to 0.96, lagging. This 

resulted in an increase in current of 4% to 7%.
25

• The voltage and current design calculations assumed the generator voltage would be 

independent of temperature. Test data confirmed the theoretical temperature coefficient 

of the magnets and demonstrated that elevated magnet temperatures can significantly 

reduce the induced EMF of the generator. The decrease in voltage required the current to 

be increased to maintain the desired power.

• The elevated current resulting from the lagging power factor and reduced EMF caused 

increased conduction losses. The increased losses resulted in increased generator 

temperatures and, consequently, increased magnet temperatures. This caused further 

reduction in the generator voltage.

The relationship between current, voltage, and magnet temperature forms a positive-feedback 

system. Adjusting the current and voltage values to account for the lagging power factor and 

thermal effects decreases the discrepancy between the design and measured values; however,

there still is some divergence between the design and measured values at greater power levels.

One of the causes of this remaining divergence is that the elevated current levels increase the 

voltage drop across the impedance of the stator circuit; this results in reduced voltage at the 

generator terminals. The decreased voltage requires more current to maintain a given power level 

which, in turn, increases the voltage drop in the stator impedance. Similar to the thermal effects 

on the magnets, this forms a positive-feedback system which exacerbates the discrepancy 

between design values and measured values. The increased voltage drop across the stator 

impedance due to elevated current is estimated to affect the magnitude of the current and voltage 

by up to 1%.

Figure 19. Generator voltage versus generator power

0

100

200

300

400

500

600

700

800

900

400 600 800 1000 1200 1400 1600

Generator Power (kW)

Voltage (V)

Design

Measured

Adjusted
26

Figure 20. Generator current versus generator power

5.1.1.2 Waveforms

The current waveforms of one quadrant of the generator at one third rated power and rated power 

are shown in Figure 21 and Figure 22. The nominal rotational speed at one third rated power is 

133 rpm and the nominal rotational speed at rated power is 191 rpm, as listed in the test matrix in

Table 10. At one third rated power, the prominence of the fifth harmonic is evident in the double 

hump of the current waveforms. As the power increases, the magnitude of the fifth harmonic—as 

well as the other higher-order harmonics—decreases and the current waveforms become more 

sinusoidal. 

0

200

400

600

800

1000

1200

1400

400 600 800 1000 1200 1400 1600

Generator Power (kW)

Current (A)

Design

Measured

Adjusted
27

-200

-150

-100

-50

0

50

100

150

200

0 0.01 0.02 0.03 0.04 0.05

Time (s)

Current (A)

Phase A

Phase B

Phase C

Figure 21. Generator current waveforms at one third rated power

-800

-600

-400

-200

0

200

400

600

800

0 0.005 0.01 0.015 0.02 0.025 0.03 0.035 0.04 0.045 0.05

Time (s)

Current (A)

Phase A

Phase B

Phase C

Figure 22. Generator current waveforms at rated power

Although the generator current waveforms become more sinusoidal as the load increases, the 

voltage waveforms become increasingly distorted as generator power increases. The voltage 

waveforms at one third rated power and rated power, in Figure 23 and Figure 24, respectively,

show large fifth and seventh harmonics. The influence of the higher-order harmonics increases 

with load.
28

-800

-600

-400

-200

0

200

400

600

0 0.005 0.01 0.015 0.02 0.025 0.03 0.035 0.04 0.045 0.05

Time (s)

Generator Voltage (V)

Phase A

Phase B

Phase C

Figure 23. Generator voltage waveforms at one third rated power

-800

-600

-400

-200

0

200

400

600

800

0 0.01 0.02 0.03 0.04 0.05

Time (s)

Generator Voltage (V)

Phase A

Phase B

Phase B

Figure 24. Generator voltage waveforms at rated power

5.1.2 Harmonics

Figure 25, Figure 26, Figure 27, and Figure 28 contain the power spectral density (PSD) of the 

generator current and voltage at one third rated power and rated power. The frequency axes of

these plots are normalized by the fundamental electrical frequency at the generator. For example, 

one corresponds with the fundamental frequency, two corresponds with the second harmonic, 

and three corresponds with the third harmonic. The fundamental electrical frequency is 52.5 Hz 

at one third rated power and 76.0 Hz at rated power. The current PSD shows a significant 

reduction in the fifth harmonic and moderate reductions in the eleventh and thirteenth harmonics
29

as the generator power is increased. The voltage PSD shows an increase in the fifth, seventh, 

eleventh, and thirteenth harmonics.

0

1

10

100

1,000

10,000

100,000

1,000,000

0 1 2 3 4 5 6 7 8 9 10 11 12 13 14

Electrical Frequency Harmonics

Power Spectral Density (A 2/Hz)

Figure 25. Generator current power spectral density at one third rated power

0

1

10

100

1,000

10,000

100,000

1,000,000

0 1 2 3 4 5 6 7 8 9 10 11 12 13 14

Electrical Frequency Harmonics

Power Spectral Density (A 2/Hz)

Figure 26. Generator current power spectral density at rated power
30

0

1

10

100

1,000

10,000

100,000

1,000,000

0 1 2 3 4 5 6 7 8 9 10 11 12 13 14

Electrical Frequency Harmonics

Power Spectral Density (V 2/Hz)

Figure 27. Generator voltage power spectral density at one third rated power

0

1

10

100

1,000

10,000

100,000

1,000,000

0 1 2 3 4 5 6 7 8 9 10 11 12 13 14

Electrical Frequency Harmonics

Power Spectral Density (V 2/Hz)

Figure 28. Generator voltage power spectral density at rated power

The total harmonic distortion (THD) of the generator current and voltage is plotted against 

generator power in Figure 29. The THD and PSD results indicate that increased load smoothes

the current waveforms and distorts the voltage waveforms. These results are consistent with 

expectations for a permanent magnet generator with a passive rectifier.
31

0%

10%

20%

30%

40%

50%

60%

0 200 400 600 800 1,000 1,200 1,400 1,600

Power (kW)

Total Harmonic Distortion

Current

Voltage

Figure 29. Total harmonic distortion of generator voltage and current

5.1.2.1 Cogging Torque

The magnetic cogging torque of the generator was measured by DRS Electric Power 

Technologies with the rotor in place and the cover removed, using the test setup shown in 

Figure 30. The torque measurement arm used a turnbuckle to rotate the rotor and a load cell 

measured the force at approximately every one eighth of a degree. The rotational position was 

measured using a scale on the rotor, as shown in Figure 31.

Figure 30. Cogging torque measurement test setup (Photo credit: Terry Lettenmaier, Global 

Energy Concepts, LLC)
32

Figure 31. Rotation scale with tics at 2-mm intervals (Photo credit: Terry Lettenmaier, Global 

Energy Concepts, LLC)

The torque measurement arm was attached to the mounting flange of the generator housing. This 

allowed the torque measurement arm to apply tangential forces to the rotor. The test setup took 

advantage of the symmetry of the housing, enabling the measurement arm to be mounted on 

either side of the housing, and increasing the effective measurement to a 52o mechanical sweep.

The torque was measured as the turnbuckle was expanded and contracted, so both clockwise 

(CW) and counterclockwise (CCW) data are included.

The torque measurements were calibrated to the position of a magnet with respect to the teeth of 

the stator. The initial position of the rotor was such that every other magnet was in D-axis 

alignment with every third stator tooth. Figure 32 contains the static cogging torque 

measurement results. In the Generator Factory Acceptance Test Report, DRS indicated the 

cogging torque period is 15o.
33 

-2.5

-2

-1.5

-1

-0.5

0

0.5

1

1.5

2

2.5

0 10 20 30 40 50 60

Generator Rotor Position (Deg)

Torque (kNm)

CW

CCW

Figure 32. Static cogging torque 

Figure 33 contains a frequency domain plot of the generator no-load torque with the generator 

speed at 133 rpm and at 169 rpm. The cogging torque periodicity indicated by DRS results in a 

cogging torque frequency of 53.2 Hz for the 133-rpm test run and 67.6 Hz for the 169-rpm test 

run. These frequencies correspond with the fundamental electrical frequency of the generator.

0.000

0.001

0.010

0.100

1.000

10.000

0 100 200 300 400 500

Frequency (Hz)

Torque (kN-m)

133 RPM

169 RPM

 

Figure 33. Generator no-load torque 
34 

5.1.2.2 Current Balance

Figure 34 shows the current imbalance between generator quadrants for a subset of the test data. 

The current imbalance between generator quadrants is relatively low and does not appear to be a 

function of load. The current imbalance between the four quadrants varies from 0.8% to 1.3% 

indicating that the core and windings are relatively uniform. The results in Figure 34 are based 

on approximately 30,000 10-second records.

0%

5%

10%

15%

20%

25%

30%

300 500 700 900 1100 1300 1500

Generator Power (kW)

Current Imbalance Between 

Quadrants (%)

Average

Min.

Max.

Std. Dev.

Figure 34. Generator current imbalance

5.1.2.3 Power Electronics

Figure 35 contains the average power factor of the inverters. The average power factor presented 

in Figure 35 is based on a total of 83 hours of data during power production. The power factor of 

the inverters is close to unity when the output power is below rated; however, the power factor 

drops below 0.99 at rated power. 
35

0.95

0.96

0.97

0.98

0.99

1.00

1.01

1.02

1.03

1.04

1.05

300 500 700 900 1100 1300 1500 1700

Total Inverter Power (kW)

Power Factor

Figure 35. Inverter power factor

5.1.3 Gearbox Characterization

5.1.3.1 Ring Gear Tooth Load Distribution

The test article planetary stage gearbox includes a unique double-helical gearing arrangement 

that inherently centers the planets in the ring gear and avoids introducing any thrust load into the 

planets. The sun gear, in turn, “floats” between the three planet gears and does not require any 

thrust reaction.

The double-helical ring gear was challenging to manufacture, but the accuracy of the gearing 

was excellent in the end. The ring gear teeth were designed with lead correction to avoid contact 

at the ends of the teeth. Each tooth was modified to have a slight crown, so that at low load the 

contact period with the mating tooth on the planet gear would be short, and would increase 

gradually as the load increases. To verify the lead correction on the ring gear teeth, strain gages 

were bonded to the root of selected teeth at four positions along the tooth from one side to the 

other. Three azimuthal locations on the ring gear were instrumented at 120° apart, and were 

designated as X, Y, and Z. Position X is 180° from top dead center (TDC). Position Y is 300°

clockwise from top dead center when looking at the rotor side of the gearbox. Position Z is 60°

clockwise from TDC when looking at the rotor side of the gearbox. For each location, the 

upwind half of the double-helical gear is designated as A and the downwind half as B. Note that 

the four strain gages at each location were configured into one bridge, and were staggered over 

11 teeth so that only one gage in the bridge would be loaded at any one time. Figure 36 shows a 

roll-out of the strain gage installation.
36

Figure 36. Ring gear strain gage configuration

Figure 37, Figure 38, and Figure 39 show typical signals from each of the strain gage bridges 

during operation at 1,104 kW (approximately two thirds full load). Note the characteristic humps 

that occur as each gage in the bridge is successively loaded.

-1.5

-1.0

-0.5

0.0

0.5

1.0

1.5

950 1000 1050 1100 1150 1200 1250

Time (ms)

Voltage (V)

 RGXA

 RGXB

Figure 37. Ring gear strain gages at X position
37

-0.4

-0.3

-0.2

-0.1

0.0

0.1

0.2

0.3

950 1000 1050 1100 1150 1200 1250

Time (ms)

Voltage (V)

 RGYA

 RGYB

Figure 38. Ring gear strain gages at Y position

-0.3

-0.2

-0.1

0.0

0.1

0.2

0.3

950 1000 1050 1100 1150 1200 1250

Time (ms)

Voltage (V)

 RGZA

 RGZB

Figure 39. Ring gear strain gages at Z position

Figure 40 shows how the peaks in the signal can be interpreted to derive the relative load at each 

gage location. This technique was applied to the data from four load stages to investigate how 

the relative strain at each gage changes with increase in load.
38

-0.3

-0.2

-0.1

0.0

0.1

0.2

950 1000 1050 1100 1150 1200 1250

Time (ms)

Voltage (V)

 RGZA

Gage 1

Gage 2

Gage 3 Gage 4

Figure 40. Measuring change in strain level

Figure 41 is plot of the tooth load distribution for each location along the teeth at the X, Y, and Z 

positions during operation near rated power.

0.1

0.3

0.5

0.7

0.9

1.1

1.3

1.5

1 2 3 4

Location Along Tooth

Relative Strain

RGYA

RGYB

RGXA

RGXB

RGZA

RGZB

Figure 41. Tooth load distribution at 1,448 kW

The load distributions across the teeth are not uniform. The potential causes for these differences 

in strain levels include the following.

• Helical cut of the gears: The expected loading distribution, as designed, is not perfectly 

uniform.
39

• Parallel alignment of planet gears: The annulus gear, due to its manufacturing processes, 

has much greater eccentricity error. The right and left helices are finished separately, 

subject to limitations of relocation after disturbing the work piece between finishing the 

two helix hands. This non-concentric location of pitch cylinders causes mating gears to 

relocate their apexes until helical thrust forces are equalized. Due to their spherical 

bearing mounting, this will be automatic for the planets. This displacement of the apex 

center at the annulus gear, in relation to the sun pinion apex, creates a slope in the planet 

axis. Not all planets have the same slope, because they are engaged at different places in 

the apex displacement pattern. Repeated test runs at various loads would quantify the 

repeatability of the strain measurements to help isolate planet gear self-alignment versus 

instrumentation or manufacturing errors.

• Axial position of the sun gear: The data in Figure 53 show changes in sun gear 

displacement as load changes, corresponding to load changes in the ring gear due to 

realignment of the planet gears.

• Input shaft alignment: The load distribution shown in Figure 41 shows the relative strain 

at position X (bottom dead center) is approximately three to six times greater than at 

positions Y and Z. This indicates the carrier has a downward vertical loading component.

This simply might be self-weight of the carrier and planet gears. It also could be due to 

uneven input-shaft loading caused by an alignment error. Vibration data from 

accelerometers on the main bearings does reveal an elevated main shaft 

once-per-revolution (1P) signal, but the accelerometers are mounted horizontally and are 

integrated circuit piezoelectric (ICP) accelerometers, which cannot reliably measure 

magnitude at frequencies below 1 Hz.

• Instrumentation errors: These could include calibration, scaling, and other errors. These 

types of instrumentation errors were not noted during or after the testing; additionally, in 

accordance with the test plan (document number N600006 R7), the ring gear strain 

measurements are relative measurements and are not calibrated to a particular load. 

• Strain gage placement: These effects could be studied using finite-element method 

models to determine whether they are significant. Placement issues can be related to the 

location and orientation of the strain gages. Location of the strain gages along the length 

of the valley between teeth is expected to have minimal effect. Location errors in the 

transverse direction (perpendicular to the valley) could be significant due to the 

asymmetrical cross-sectional shape of the tooth, but tip clearances of the planet gear tooth 

tip to the ring gear root preclude much alignment error before the strain gage and wiring 

would interfere with the mesh. Angular errors in gage placement also could be 

significant, but these were not studied.

Figure 42 and Figure 43 show the measured load distribution for the gages at the Z location, 

derived by measuring the relative change in strain using the method depicted above. Note that, as 

the load increases from 573 kW to 1,448 kW, the line tends to level out. This indicates that the 

load is more evenly distributed across the tooth. This change in load distribution possibly is due 

to the helical cut of the gears. As the load increases, the gear teeth deflect and transfer the load 

more equally along the tooth face.
40

0.1

0.12

0.14

0.16

0.18

0.2

0.22

0.24

0.26

1 2 3 4

Location Along Tooth

Relative Strain

Power = 573 kW

Power = 814 kW

Power = 1266 kW

Power = 1448 kW

Figure 42. Tooth load distribution from the strain gauge bridge, RGZA

0.1

0.12

0.14

0.16

0.18

0.2

0.22

0.24

1 2 3 4

Location Along Tooth

Relative Strain

Power = 573 kW

Power = 814 kW

Power = 1266 kW

Power = 1448 kW

Figure 43. Tooth load distribution from the strain gauge bridge, RGZB

Figure 44 and Figure 45 show the measured load distribution for the gages at the Y location. The 

signals from the gages at the Y location show a more pronounced flattening out with load 

increase. The data also show some planet misalignment between the A and B meshes, which 

decreases with load as expected in the design. Figure 45 shows an undesirable load distribution 
41

at low power, indicating the planets might not self-align as freely at lesser loads. The criticality 

of this condition on planet bearing roller skidding and gear-tooth loading should be investigated

further. Longer runs at various power settings are necessary to determine the duration and 

magnitude of misalignment, and strain gage load calibration would be necessary to compare the 

tooth load with the design-load limits.

0.1

0.12

0.14

0.16

0.18

0.2

0.22

0.24

0.26

1 2 3 4

Location Along Tooth

Relative Strain

Power = 573 kW

Power = 814 kW

Power = 1266 kW

Power = 1448 kW

Figure 44. Tooth load distribution from the strain gauge bridge, RGYA
42

0.1

0.12

0.14

0.16

0.18

0.2

0.22

0.24

0.26

1 2 3 4

Location Along Tooth

Relative Strain

Power = 573 kW

Power = 814 kW

Power = 1266 kW

Power = 1448 kW

Figure 45. Tooth load distribution from the strain gauge bridge, RGYB

Figure 46 and Figure 47 show the measured load distribution for the gages at the X location. The 

relative strains at location X are several times greater than at locations Y and Z for each location 

along the tooth.

0.1

0.3

0.5

0.7

0.9

1.1

1.3

1.5

1 2 3 4

Location Along Tooth

Relative Strain

Power = 573 kW

Power = 814 kW

Power = 1266 kW

Power = 1448 kW

Figure 46. Tooth load distribution from the strain gauge bridge, RGXA
43

Figure 47. Tooth load distribution from the strain gauge bridge, RGXB

Although the general shape of the curves for location X exhibit good behavior and are similar in 

shape to those for the other locations, the gages at the X location show very little change in 

distribution across the length of the tooth as load increases. This could be due to the increased 

loading from self-weight or the alignment of the input shaft. The loading on the planet bearing 

rollers is distributed more evenly across the rollers (as described in Section 5.1.3.5), 

consequently there is much less flattening in relative strain from position 1 to 4 as load increases. 

The tooth loading due to sun pinion offset corresponds to the design documentation, and is 

shown in Figure 48.

Figure 48. Tooth loading caused by pinion offset

A4 B1

0.1

0.3

0.5

0.7

0.9

1.1

1.3

1.5

1 2 3 4

Location Along Tooth

Relative Strain

Power = 573 kW

Power = 814 kW

Power = 1266 kW

Power = 1448 kW
44

5.1.3.2 Sun Gear and Torque Tube Motion

The gearbox arrangement includes a “floating” sun gear that is captured by the three planet gears

on one end, and which transmits torque to the generator through a hollow torque tube that is 

fixed to the generator rotor as shown in Figure 49. The sun gear connects to the torque tube with 

a splined connection, so it can move a small amount axially. Additionally, the torque tube is 

connected to the generator rotor through a flexible diaphragm so that it can compensate for some 

axial misalignment. The intent of these features is to isolate the sun gear and the generator so that 

only torque is transmitted, and any misalignment of the various components will not be 

transferred into axial misalignment of the sun gear and planet mesh.

Figure 49. Test article cutaway

To verify the effectiveness of the splined connection between the sun gear and torque tube, 

proximity sensors were installed to measure the axial displacement of the sun gear and of the 

torque tube. Figure 50 shows the axial motion of the sun gear and of the torque tube when the 

generator output was 814 kW. The planet passage trace in Figure 50 indicates the periodicity of 

the planet gears. The frequency of the planet gear passage is three times the frequency of the 

carrier rotation. Each spike in the planet passage trace marks the passage of one of the planet 

gears past a fixed point on the ring gear. It is expected that because the sun gear is captured by 

the planet gears—which in turn follow the circumferential path of the ring gear apex—any 

variation in the ring gear apex due to manufacturing error or distortion will be reflected in the 
45

sun gear axial motion. The traces show that the axial motion of the torque tube is on the order of 

150 microns peak-to-peak. Even without the subtraction of sun pinion displacement (spline slip),

as shown in Figure 50, this is well within the design parameters, which provide calculations for 

sun pinion displacements of up to 550 microns.

-0.10

-0.08

-0.06

-0.04

-0.02

0.00

0.02

0.04

0.06

0.08

0.10

0 1 2 3 4 5

Time (s)

Displacement (mm)

Torque tube axial displacement

Sun gear axial displacement

Planet passage

Figure 50. Sun gear and torque tube axial displacement

It also is interesting to evaluate the effectiveness of the spline connection in isolating the ring 

gear and the torque tube. Figure 51 shows the relative axial motion between the sun gear and the 

torque tube, again with the planet gear passage trace included. Note that the relative motion is 

almost equal to the sun pinion motion, indicating that the spline is not compensating for gear 

mesh thrust loading to nearly the extent that the flexible diaphragm of the torque tube

compensates.
46

-0.10

-0.08

-0.06

-0.04

-0.02

0.00

0.02

0.04

0.06

0.08

0.10

0 1 2 3 4 5

Time (s)

Displacement (mm)

Relative displacement

Planet passage

Figure 51. Relative sun gear and torque tube axial displacement

Lastly, a proximity sensor was installed to measure the radial motion of the torque tube. This 

measurement, shown in Figure 52, reflects the misalignment between the axis of the sun gear and 

the axis of the generator or ring gear. The frequency of the displacement identifies the pertinent 

misalignment. Superposed is the marker for the generator 1-per-revolution (generator 

revolution). Note that the range of motion is approximately 200 microns, which is greater than 

expected, but it could be a reasonable misalignment for the gearbox and generator housing which 

support these large parts.
47

-0.15

-0.10

-0.05

0.00

0.05

0.10

0.15

0 1 2 3 4 5 6 7 8 9 10

Time (s)

Displacement (mm)

Torque tube radial displacement

Generator Revolution

Figure 52. Torque tube radial displacements

Figure 53, Figure 54, Figure 55, and Figure 56 are plots of the mean, minimum, and maximum 

sun gear axial displacement; torque tube axial displacement; torque tube radial displacement; and 

relative axial displacement between the sun gear and the torque tube. Figure 55 shows that radial 

position does not change with load. This might further support other data that indicate an 

assembly alignment error.

300 400 500 600 700 800

0

0.2

0.4

0.6

0.8

1

Main shaft torque (kNm)

Position (mm)

Mean

Min

Max

Figure 53. Sun gear axial position
48

300 400 500 600 700 800

2

2.2

2.4

2.6

2.8

3

Main shaft torque (kNm)

Position (mm)

Mean

Min

Max

Figure 54. Torque tube axial position

300 400 500 600 700 800

1.8

2

2.2

2.4

2.6

2.8

Main shaft torque (kNm)

Position (mm)

Mean

Min

Max

Figure 55. Torque tube radial position

300 400 500 600 700 800

1.4

1.6

1.8

2

2.2

Main shaft torque (kNm)

Position (mm)

Mean

Min

Max

Figure 56. Relative axial displacement between sun gear and torque tube
49

Figure 57 is an overlay plot of the sun gear, torque tube, and relative sun gear–torque tube 

displacements. Although the position of the torque tube changes with load, the magnitude of 

displacement of the sun gear and torque tube is independent of load.

300 350 400 450 500 550 600 650 700 750 800

-0.1

-0.05

0

0.05

0.1

0.15

Main shaft torque (kNm)

Displacement (mm)

SGD

TTA

TTR

Relative SGD-TTA

Figure 57. Sun gear and torque tube axial displacements

5.1.3.3 Planet Gear Instrumentation

Instrumentation was installed in the planetary section of the gearbox to measure various 

characteristics of the planet gear and bearing loading. Figure 58 shows the location of this 

instrumentation (discussed in more detail below). Figure 59 shows the instrumentation installed.
50

Planet Bearing Load @ 

Top-dead-center

Planet Bearing Load @ 72 degrees off TDC

Planet Gear Axial Displacement sensor

Magic Roller

Figure 58. Planetary gear instrumentation location

Figure 59. Planetary gear installed instrumentation. 

(Photo credit: Chris Walford, DNV Renewables)
51

5.1.3.4 Planet Gear Alignment

As noted, the double-helical gearing arrangement is intended to capture the planet gears axially. 

Any axial motion of the planet gears is a reflection of the axial run-out of the ring gear apex. To 

measure this motion and detect axial deflection of one planet gear, proximity sensors were 

installed on the carrier at four different locations relative to the planet carrier—outboard, leading, 

lagging, and inboard. The inboard sensor was not functional and is excluded from the results. 

The input shaft does not impart thrust loads during this test, therefore the carrier can be assumed 

to be axially unloaded during testing, making it a suitable reference point for axial-deflection 

measurements relative to the housing.

Figure 60 shows a representative time series of the axial displacement from the sensor at the 

lagging position. Note that a very small component of the generator 1P motion is reflected in the 

planet motion (approximately 5 to 10 microns) through the torque tube and sun gear, but this is 

overwhelmed by the 1P of the carrier. The axial deflections of this planet are shown to vary by 

only approximately 25 microns. There appears to be relatively little ring gear axial variation as 

the planet travels around the ring gear perimeter, assuming that the carrier axis is aligned closely 

with the ring gear axis.

Figure 50, however, shows the sun gear axial deflections to vary approximately 150 microns due 

to planet passage. This is significantly greater than the axial variations of the single, 

instrumented planet as measured relative to the carrier. This is consistent with the design 

expectations for the floating sun pinion.

DNV did not perform additional testing to investigate the causes for these findings. The effects 

of generator alignment—or misalignment—relative to the ring gear, planets, and sun gear on 

tooth loading were not investigated. Misalignment between the generator and the gear box, 

however, could be the cause for some of the results.
52

0.380

0.385

0.390

0.395

0.400

0.405

0.410

0.415

0.420

0 1 2 3 4 5 6 7 8

Time (s)

Displacement (mm)

Planet gear displacement, lagging

Generator revolution

Figure 60. Planet gear axial displacement, lagging

The outboard, leading, and lagging axial position of the instrumented planet gear position is 

shown in Figure 61, Figure 62, and Figure 63 as a function of load. The inboard axial position is 

not presented because the inboard displacement sensor was not functional. For each of these 

locations, the axial deflection scatter about the mean value frequently was less than 30 microns, 

indicating low sensitivity in planet axial displacement to applied torque as measured at the 

carrier.

300 400 500 600 700 800

3.02

3.03

3.04

3.05

3.06

Main shaft torque (kNm)

Position (mm)

Mean

Min

Max

Figure 61. Planet gear axial displacement, outboard
53

300 400 500 600 700 800

3.03

3.04

3.05

3.06

Main shaft torque (kNm)

Position (mm)

Mean

Min

Max

Figure 62. Planet gear axial displacement, leading

300 400 500 600 700 800

0.38

0.39

0.4

0.41

0.42

Main shaft torque (kNm)

Position (mm)

Mean

Min

Max

Figure 63. Planet gear axial displacement, lagging

5.1.3.5 Planet Bearing Load Distribution

The planet bearing load signals PBL1 through PBL8 derive from strain gages bonded to axial 

slots machined into the inner race of one of the planet CARB bearings. The strain gages react to 

the contact pressure from the roller load as the rollers pass over the slots. Figure 64 shows side 

and section views of the planet bearing. The orange outboard arrows indicate the load applied 

from the sun and ring gear during operation. The blue arrow indicates the reaction load in the 

bearing. The section view shows the unique “barrel” shape of the CARB rollers. This shape 

allows for some angular and axial misalignment between the inner and outer races of the bearing 

to accommodate inaccuracies and deflections in the planet carrier and gearing.
54 

Figure 64. Planet bearing, side and section view

Figure 65 is a photograph of the instrumented planet bearing. The strain gage wires can be seen 

exiting the grooves in the bearing inner race (filled with black bonding compound). 

Figure 65. Instrumented planet bearing (Photo credit: Chris Walford, DNV Renewables Inc.)

Figure 66 shows a side view of the inner race of the planet bearing, and a “rollout” of the strain 

gage placement within each groove. There are eight strain gages; five in Slot A and three in Slot 

B. Slot A is oriented so that it is TDC relative to the reaction load in the bearing; this slot sees 

the greatest roller loads. Slot B is oriented 72° off from TDC. The roller loads at this slot 

location are much less than those found at Slot A. The reason for using two slots is to estimate 

the relative load at each azimuth location, and thus estimate the roller load distribution within the 

bearing.
55

Figure 66. Planet bearing inner race, side view

Originally, the intent was to calibrate the planet bearing load signals to the applied load so that 

actual roller loads could be measured during operation. A calibration fixture was fabricated to 

enable measured loads to be applied to the instrumented bearing. Figure 67 shows the calibration 

fixture. The planet bearing inner race is mounted to an axle which is the same size as the planet 

carrier pin. This axle is mounted into roller bearings on either end so that it is free to rotate. The 

planet bearing outer race is mounted into a loading block that is fabricated to apply load at either 

side of the bearing (TDC +/- 90°) to simulate the actual loading of the planet as shown in 

Figure 64 (above).
56

Figure 67. Bearing load calibration fixture (Photo credit: Chris Walford, DNV Renewables Inc.)

The load was applied at staged levels by tightening the threaded tie-rods and measuring at the 

load cell. While the load was applied, the inner race was rotated approximately 180° to move one 

roller through the TDC position. The orientation of the inner race was recorded with a rotary 

potentiometer. Figure 68 shows an example of the measurement taken from the center gage in 

Slot A (PBL3) for the series of applied loads. Note that, in general, the traces are symmetric, 

repeatable, and increasing approximately linearly with applied load.
57

-0.45

-0.40

-0.35

-0.30

-0.25

-0.20

-0.15

-0.10

-0.05

71 72 73 74 75 76 77 78 79 80 81 82 83 84 85 86 87 88 89 90

Absolute Angle (deg)

Signal (V)

100 kip

50 kip

25 kip

125 kip

135 kip

1kip

Figure 68. CARB bearing calibration, PBL3 (Slot A)

Note that the calibration procedure only measured the gage response for one roller. This was 

based on the assumption that because all the rollers would be identical, the response to any single 

roller would be the same. Unfortunately, upon review of the measurements recorded on the 

actual test article, it became obvious that this assumption was erroneous. Figure 69 shows a time 

series recorded with the drivetrain operating at 1,100 kW (approximately two thirds of rated 

power) for signal PBL2. This is the gage located near the center of Slot A at TDC. The time 

segment covers the passage of a complete roller set (20 rollers) over the slot. The amplitude is 

not calibrated to the load, but it has been normalized to the average signal value to remove the 

signal offset from the strain gage.

-0.15

-0.10

-0.05

0.00

0.05

0.10

0.15

0.20

0 500 1000 1500 2000

Time (ms)

Non-dimensional

Figure 69. Planet bearing load (PBL2) at 1,100 kW
58

Note that the signal peaks occur in groups of four equally spaced intervals, repeated five times 

for the 20-roller cycle. This pattern is due the bearing cage having five structural supports that 

space the rollers into five groups, as depicted in Figure 65 and visible in Figure 67. As shown in 

Figure 69, the rollers adjacent to the gap receive greater loading (two adjacent tall peaks) than do 

the rollers in the middle of the grouping (two lower peaks). Also note that the amplitude varies 

from roller to roller, as shown in Figure 70 for the next sequential set of signals. This next set is

nearly identical to the first set in that the general pattern is repeated but the magnitudes are 

different. This difference probably is due to slight variations in roller diameter or profile, 

although this has not been confirmed with physical measurements.

-0.15

-0.10

-0.05

0.00

0.05

0.10

0.15

0.20

2350 2850 3350 3850 4350

Time (ms)

Non-dimensional

Figure 70. Planet bearing load (PBL2) at 1,100 kW

This variation in signal due to the asymmetrical bearing spacing rendered the calibration effort 

useless (at least for any degree of accuracy) because it is not possible to assign a specific roller to 

a specific signal in the sequence. Nevertheless, the data still reveal important information 

regarding the load distribution among the rollers—and the load distribution along any one 

roller—as the applied load changes. Figure 71 shows the change in average peak signal with 

drivetrain power level for the gages at Slot A (located at TDC). The signals for gages PBL1 and 

PBL5—each located at the outer edge of the slot—have been averaged to remove any offset due 

to axial misalignment of the inner and outer bearing races. The same rationale applies to PBL2 

and PBL4, which are located on either side of the center of the slot. Note that the signals increase 

approximately linearly with applied load.
59

0.00

0.01

0.02

0.03

0.04

0.05

0.06

400 600 800 1000 1200 1400 1600

Drive Train Input Power (kW)

Non-dimensional Signal

PBL1 & PBL5 avg (Slot A edge)

PBL2 & PBL4 avg (Slot A center)

Figure 71. Planet bearing strain gage peak loads at slot A

Figure 72 shows the same data for the gages in Slot B, located at 72° off of TDC. The signals in 

this slot increase at a faster rate than the do signals in the TDC slot. This shows that the roller 

load zone spreads out over more rollers as the applied load increases. This phenomenon is seen 

clearly by examining the ratio of the signals at Slot B to the signals at Slot A, as shown in Figure 

73. At the end of the load zone, the load that is borne by the rollers increases dramatically, from 

27% to nearly 70% of that of the rollers located at TDC, which bear the greatest load.

0.00

0.01

0.02

0.03

0.04

0.05

0.06

400 600 800 1000 1200 1400 1600

Drive Train Input Power (kW)

Non-dimensional Signal

PBL6 & PBL8 avg (Slot B edge)

PBL7 (Slot B center)

Figure 72. Planet bearing strain gage peak loads at slot B
60

0%

10%

20%

30%

40%

50%

60%

70%

80%

90%

100%

400 600 800 1000 1200 1400 1600

Drive Train Input Power (kW)

Non-dimensional Signal

PBL6/PBL2 (Slot B/Slot A)

Figure 73. Planet bearing strain gage peak loads, ratio of slot B to slot A

The effect of the roller load zone “spreading out” over more rollers as applied load increases is 

expected for a roller bearing, but the unique loading of a planet gear greatly increases the effect. 

The unusual bearing load zone of a gear that is loaded by teeth 180° apart can be seen in Figure 

74. The image on the left in Figure 74 shows the theoretical roller load distribution for a bearing 

that is loaded at one point, as with a parallel-shaft arrangement. The image on the left is derived 

from a finite-element analysis of the planet gear used in this test article (executed by SKF 

Industries, the manufacturer of the CARB bearing). The arrows show the relative magnitude of 

the load on each roller with each type of gear. Note that the load zone for the planet gear is 

distributed more uniformly and the maximum roller load is less than for the rigid gear. The red 

arrow indicates the measured relative roller load at a planet load of 516 kN (approx 1,450 kW 

input power), compared with the predicted roller load for a planet load of 600 kN.
61

Figure 74. Roller load distribution

Another interesting finding from the planet bearing load data is an estimate of the pressure 

distribution along the length of the roller. The CARB roller is not a true cylinder; the pressure 

under very light loads therefore is localized under the middle of the roller. The pressure 

distribution of heavy loads was expected to spread out over the length of the roller. Figure 75

shows that effect, with the ratio of the signal at the gages toward the edges of the slots to the 

signals toward the center of the slots. For both Slot A and Slot B this ratio increases with load.

Figure 75. Planet bearing load distribution

0%

10%

20%

30%

40%

50%

60%

70%

80%

90%

100%

400 600 800 1000 1200 1400 1600

Drive Train Input Power (kW)

Non-dimensional Signal

PBL1/PBL2 at Slot A (TDC)

PBL6/PBL7 at Slot B (72 deg)
62

5.1.3.6 Magic Roller

One of the planet bearings was instrumented by the manufacturer, SKF Group, using its 

proprietary “Magic Roller” sensor. Strain gages are installed inside one of the rollers. The gages 

react to roller load as the roller moves through the active load zone for the bearing. The signal 

from the gages is conditioned in the rotating frame and then transmitted to a stationary receiver. 

Figure 76 shows a typical signal from the Magic Roller. Note the periodic variation as the strain

gages in the roller are loaded and unloaded at approximately 5 Hz. This signal is modulated by 

the roller moving through the load zone approximately every two seconds.

-2.0

-1.5

-1.0

-0.5

0.0

0.5

0 2 4 6 8 10

Time (s)

Volts

Figure 76. Magic Roller waveform (generator power = 1,457 kW)

The signal from the Magic Roller was processed by SKF to derive the bearing roller load zone 

for several operating points. Figure 76 shows the measured mean bearing load for operation at 

1,457 kW, and 191 rpm. Due to a communication error, the calibration is off by a factor of 2.2. 

The correct value of 517 kN (235 ∙ 2.2) compares well with the 519 kN theoretical load.

2

 2 Torque = 1,457 kW ∙ (60/2pi) / (191/9.3) rpm = 677.5 kN-m. With a center distance of 435 mm, the radial load for 

each planet is 677.2 kN-m / (3 ∙ 0.435) = 519 kN.
63

1 1.5 2 2.5 3 3.5 4

0

50

100

150

200

250

300

350

Load case 1457kW, mean calculated radial load 235 kN

Revolutions

Load, kN

Figure 77. Magic Roller mean load (generator power = 1,457 kW)

Figure 78 shows the measured roller load zone for operation at 1,457 kW. Again, with the 

scaling corrected by the factor of 2.2, the maximum roller load is 99 kN (note that the SKF 

source data were not provided to modify charts). The shape of the roller load distribution, 

indicated by the red line, agrees well with the distribution predicted by finite element analysis in

Figure 74 above.

 0.5

 1

 1.5

30

210

60

240

90

270

120

300

150

330

180 0

Bearing loaded zone. Max Roller load 45 kN

Figure 78. Magic Roller load zone (generator power = 1,457 kW)
64

5.1.3.7 Transient Response

The transient response of the drivetrain was tested by commanding step changes in power to the 

inverters while the dynamometer was operating in constant-speed mode. Figure 79 and Figure 80

show the torque response of the drivetrain when coupled to the dynamometer. A quantitative 

analysis of the transient response has not been performed because of the difficulty in separating 

the response of the drivetrain from the response of dynamometer.

Figure 79. Torque response run #1 (rotor speed = 15.5 rpm)

Figure 80. Torque response run #2 (rotor speed = 15.5 rpm)

0

50

100

150

200

250

300

350

400

0 1 2 3 4 5 6 7 8 9 10

Time (s)

Torque (kNm)

0

75

150

225

300

375

450

525

600

Power (kW)

Generator torque (kNm)

Rotor torque (kNm)

Generator power (kW)

0

50

100

150

200

250

300

350

400

0 1 2 3 4 5 6 7 8 9 10 11 12 13

Time (s)

Torque (kNm)

0

75

150

225

300

375

450

525

600

Power (kW)

Generator torque (kNm)

Rotor torque (kNm)

Generator power (kW)
65

5.2 Thermal Performance

5.2.1 Gearbox Fluid Cooling System

Throughout the testing of the drivetrain, the temperature of the gearbox lubricant remained 

significantly below the maximum allowable temperature of 80°C. For a majority of the duration 

of generator testing, the radiator fan of the gearbox fluid cooling system was never turned on. 

During the testing shown in Figure 81, the radiator fan was not turned on until approximately 

1:30 p.m., when the lubrication temperature reached 59°C. This is the maximum temperature 

attained by the lubrication throughout all of the drivetrain testing. With the fan on, the 

lubrication temperature dropped sharply. Due to the great efficiency of the gearbox, the demand 

on the gearbox fluid cooling system was small. The heat loss in the gearbox was evacuated 

effectively by the cooling system. For the test run shown in Figure 81, the average lubrication

supply pressure is 22.5 psi and average lube flow rate is 79.7 liters per minute.

30

35

40

45

50

55

60

65

6:30 AM7:30 AM8:30 AM9:30 AM

10:30

AM

11:30

AM

12:30

PM

1:30 PM2:30 PM

Time (s)

Temperature (deg C)

200

400

600

800

1,000

1,200

1,400

1,600

Power (kW)

Lube Return Temperature

Lube Supply Temperature

Generator Power

Radiator fan

turned on

Figure 81. Gearbox lubrication system temperature

5.2.2 Generator Stator Fluid Cooling System

Although very few thermal performance data were gathered for the parallel winding 

configuration, enough data exist to compare the winding temperatures of the parallel 

configuration to those of the series configuration at two power levels. Figure 82 and Figure 83

compare the average winding temperatures at 600 kW and 700 kW, respectively, for the parallel 

and series winding configurations. The winding temperatures in Figure 82 and Figure 83 were 

measured at the outside of the winding end-turn for both the parallel and series winding 

configurations. Thermistors were used to measure winding temperature in the parallel 

configuration, and RTDs were used in the series configuration.
66

There is a significant difference between the temperatures of the parallel and series winding 

configurations. In the 600-kW test runs (Figure 82), the series winding and parallel winding 

temperatures are increasing slightly. The rate of change in temperature for both winding 

configurations is 0.03°C/min over the 45-minute set of data presented. The series winding 

temperature is more than 20°C hotter than the parallel winding temperature. Although the 

ambient interior temperature of the DTF is not available for both test runs, the ambient outside 

temperature for both test runs was near 30°C.

In the 700-kW test runs shown in Figure 83, the parallel winding temperature increases slightly 

and the series winding temperature decreases slightly. Both temperatures, however, appear to be 

approaching steady-state. The rate of change of the parallel winding temperature is 0.03°C/min 

and the rate of change of the series winding temperature is -0.02°C/min. It is unlikely that the 

35°C gap between the winding temperatures would decrease by more than 5°C to 10°C. This 

leaves a discrepancy of at least 25°C to 30°C. As noted above, the ambient interior temperature 

of the DTF is unavailable; however, the ambient outside air temperature during the parallel 

winding test run was approximately 5°C hotter than during the series winding test run. It is likely 

that the winding temperature discrepancy in Figure 83 would be even greater if the ambient 

temperature was the same for both of the test runs. The cause of the great discrepancy between 

the temperatures of series and parallel windings is unknown. The magnitude of the discrepancy 

points to a design flaw in the series windings.

Figure 82. Winding temperature comparison at 600 kW

20

30

40

50

60

70

80

90

100

110

120

0 5 10 15 20 25 30 35 40 45

Time (min)

Temperature (deg C)

Average Parallel Winding 

Average Series Winding
67

Figure 83. Winding temperature comparison at 700 kW

Although the drivetrain was able to operate at rated power for brief periods, 1.0 MW was the 

greatest power level at which the winding temperatures of the generator stabilized below the 

design limits. Figure 84 is a plot of the temperature of the inside of the winding end-turn, and the 

temperature of coolant supply and return. Throughout the testing of the drivetrain, the temperature inside the winding end-turn was significantly hotter than any of the other measured stator 

temperatures. At 1.0 MW, the winding temperatures stabilized at slightly less than the 165°C 

design limit, and the coolant temperatures stabilized well below the 100°C limit.

20

30

40

50

60

70

80

90

100

110

120

0 5 10 15 20 25 30 35 40 45

Time (min)

Temperature (deg C)

Average Parallel Winding 

Average Series Winding 
68

Figure 84. Temperature run at 1.0 MW

At 1.2 MW, the winding end-turn inside temperature exceeds the 165°C limit, as shown in 

Figure 85. Although the generator temperatures were beginning to stabilize in the test run shown 

in Figure 85, the test was halted to prevent damaging the windings and magnets. As shown in 

Figure 84, the coolant temperatures were hotter in the 1.0-MW test run than in the 1.2-MW test 

run, because the ambient temperature of the DTF was significantly greater. At 1.2 MW, the 

coolant temperatures still are significantly less than the design limits.

Generator Power = 1.0 MW

20

40

60

80

100

120

140

160

180

200

12:28 PM12:36 PM12:43 PM12:50 PM12:57 PM

1:04 PM1:12 PM1:19 PM1:26 PM1:33 PM1:40 PM1:48 PM1:55 PM2:02 PM2:09 PM

Time

Temperature (deg C)

Winding Inside of End-Turn

Coolant Return

Coolant Supply
69

Figure 85. Temperature run at 1.2 MW

At rated power, the winding temperature exceeds the design limit within a few minutes. The

rated power test run shown in Figure 86 was performed prior to the installation of RTDs at 

several locations on the stator. Only the outside of the winding end-turn temperature was 

measured during this rated power run. Based on experimentation, the outside of the end-turn 

typically was 20°C cooler than the inside of the end-turn. The windings therefore probably

exceeded the design limit when the temperature outside of the end-turn reached approximately 

145°C. Although the windings are significantly overheated at the end of this rated power run, the 

coolant temperatures are nearly 50°C less than the design limit.

Generator Power = 1.2 MW

20

40

60

80

100

120

140

160

180

9:00 AM9:07 AM9:14 AM9:21 AM9:28 AM9:36 AM9:43 AM9:50 AM9:57 AM

10:04 AM10:12 AM10:19 AM10:26 AM

Time

Temperature (deg C)

Winding Inside of End-Turn

Coolant Return

Coolant Supply

Radiator fan 

turned on

Ambient temp. 

decreased
70

Figure 86. Temperature run at 1.5 MW

The primary impetus for measuring temperatures at several stator locations was to determine 

what caused the stator windings to overheat when the coolant temperatures were relatively low.

The designed flow of heat from the stator windings to the coolant fluid (Glycol) is shown in 

Figure 87. The heat of the stator windings is conducted through the core and back iron to the 

cold plate and then from the cold plate to the Glycol. The maximum design temperature of the 

windings is 165°C at rated power. The designed temperature drop from the windings to the cold 

plate is 85°C, which results in a maximum cold plate temperature of 80°C. The sets of RTDs 

installed on the stator enabled verification of each of the temperature drops shown in Figure 87.

Generator Power = 1.5 MW

20

40

60

80

100

120

140

160

180

6:51 AM 6:54 AM 6:56 AM 6:59 AM 7:02 AM 7:05 AM 7:08 AM 7:11 AM

Time

Temperature (deg C)

Winding Outside of End-turn

Coolant Return

Coolant Supply

Estimated over 

temperature point
71

Figure 87. Design stator temperature breakdown

The greatest power level for which the drivetrain was able to reach thermal equilibrium was 

1.0 MW. At levels greater than 1.0 MW, the winding and magnet temperatures approached or 

exceeded the design limits. The temperature data in Figure 88, Figure 89, Figure 90, and 

Figure 91 were gathered after thermal equilibrium was reached for a 1.0-MW test run. Figure 88

is a plot of the measured stator temperatures at 1.0 MW. The segment temperatures are relatively 

consistent, indicating an even thermal distribution around the stator.

Figure 88. Stator temperature breakdown at 1.0 MW

0

20

40

60

80

100

120

140

160

180

200

Cold plate Back iron Laminated

tooth

Inside of endturnOutside of

end-turn

Stator Location

Temperature Drops (deg C)

Segment 2

Segment 4

Segment 9

Segment 11
72

Figure 89 is a plot of the measured stator temperatures at 1.0 MW averaged over the four segments, as well as the design stator temperatures at 1.5 MW. At 1.0 MW, the average temperature 

inside of the winding end-turn and the laminate tooth temperatures both were within 7°C of the 

rated-power design values. The conduction losses in the stator windings were proportional to the 

square of the current, therefore the conduction loss at 1.5 MW is approximately 225% of the loss 

at 1.0 MW. The 7°C margin between the measured design values at 1.0 MW and 1.5 MW is not 

sufficient to accommodate increased loss at 1.5 MW.

Figure 89. Stator temperature comparison

Figure 90 shows the measured temperature differentials between various elements of the stator. 

The temperature differentials are consistent across the four generator segments, indicating even 

heat flow from the windings to the coolant system.

0

20

40

60

80

100

120

140

160

180

Cold plate Back iron Laminated tooth Inside of endturnOutside of endturn

Stator Location

Temperature (deg C)

Measured Temps. at 1.0 MW

Design Temps. at 1.5 MW
73

Figure 90. Stator temperature drops at 1.0 MW

Figure 91 contains the measured stator temperature differentials at 1.0 MW averaged over the 

four segments, as well as the design stator temperature differentials at 1.5 MW.

Figure 91. Stator temperature drop comparison

The two potential causes of the overheating of the generator when running at power levels 

greater than 1.0 MW are the greater than expected losses in the generator, and the greater than 

expected thermal resistance between the windings and the cold plate. The cold plate and 

generator coolant temperatures remained low throughout testing. It therefore is unlikely that the 

0

10

20

30

40

50

60

70

80

90

100

110

Cold plate TO

backiron

Backiron TO

tooth

Tooth TO

inside of endturn

Tooth TO

outside of

end-turn

Cold plate TO

max winding

Stator Location

Temperature Drops (deg C)

Segment 2

Segment 4

Segment 9

Segment 11

0

20

40

60

80

100

120

Cold plate TO

backiron

Backiron TO

tooth

Tooth TO inside

of end-turn

Tooth TO

outside of endturn

Cold plate TO

winding inside

of end-turn

Stator Location

Temperature Drops (deg C)

Measured Drops at 1.0 MW

Design Drops at 1.5 MW
74

elevated temperatures resulted from an undersized cooling system. The winding temperatures 

increased dramatically from the parallel winding configuration to the series winding 

configuration, therefore a greater than expected loss in the windings seems to be the most 

probable cause of the poor thermal performance.

5.2.3 Generator Rotor Air / Water Cooling System

As described, the temperature of the magnets was measured using temperature probes and 

thermal imaging, and calculated based on the open-circuit voltage of the generator. Figure 92 is a 

plot of the magnet temperature calculated based on the open-circuit voltage after select test runs. 

Thermal equilibrium was not reached in the test runs shown in Figure 92. Although the 

performance of the magnets is hindered by increased temperature, the degradation of the 

performance temporarily is less than the specific temperature threshold. The magnets suffer 

permanent degradation at temperatures that are greater than 100°C to 110°C. During testing the 

magnet temperatures approached this temperature limit. The most likely cause of the high 

magnet temperatures was the high winding temperatures.

Figure 92. Magnet temperature

5.3 Efficiency

The efficiency of the drivetrain was measured using both the input-output method and the 

calorimetric method. The efficiency plots in this section show the efficiency predicted during the 

design and the efficiency measured during testing. It also includes design efficiency values that 

were adjusted based on the elevated current measured during testing. The intent of including the 

adjusted design values is to illustrate the theoretical impact of the elevated current and to validate 

the measured efficiencies.

0

10

20

30

40

50

60

70

80

90

100

400 600 800 1000 1200 1400 1600

Generator Power (kW)

Magnet Temperature (°C)
75

5.3.1 Input-Output Method

The input-output efficiency was calculated using measurements from the torque spool of the 

dynamometer, and using the electrical power measurements at the terminals of the generator and 

the connection point to the grid. Figure 93 is a plot of the combined gearbox and generator 

efficiency broken down by day. The efficiency values in this plot were calculated using the 

dynamometer low-speed shaft torque and speed measurement, and the electrical power 

measurement at the output of the generator. The data in Figure 93 are from four days of testing 

(August 7–10, 2007). The temperature dependence of the generator voltage causes the efficiency 

of the drivetrain to vary slightly, depending on the operating conditions. The extreme daily 

variability shown in Figure 93, however, points to malfunctioning measurement equipment. The 

data in Figure 93 for 400 kW show a variation in efficiency of 91% on August 10 and 98% on 

August 7. These results indicate that the accuracy and precision of the input and output power 

measurements are insufficient to perform efficiency calculations. The low-speed torque 

measurement likely is the greatest source of error or uncertainty in the efficiency calculations.

Figure 93. Combined gearbox and generator input-output efficiency by day

Figure 94 is a plot of the combined gearbox and generator input-output efficiency for the test 

runs when the drivetrain approached or achieved thermal equilibrium. Adjustment of the design 

efficiency calculations to include the measured current decreases the generator efficiency by 

almost 0.5% above 1.0 MW. It is difficult to evaluate the performance of the drivetrain prototype 

based on the data in Figure 94 because of the variability in efficiency (shown in Figure 93).

89%

90%

91%

92%

93%

94%

95%

96%

97%

98%

99%

0 200 400 600 800 1,000 1,200 1,400 1,600 1,800

Power (kW)

Efficiency (%)

8/7

8/8

8/9

8/10
76

Figure 94. Combined gearbox and generator input-output efficiency

Figure 95 is a plot of the input-output efficiency of the power electronics and transformers for

the test runs when the drivetrain approached or achieved thermal equilibrium.

Figure 95. Power electronics and transformer input-output efficiency

90%

91%

92%

93%

94%

95%

96%

97%

98%

99%

100%

0 200 400 600 800 1000 1200 1400 1600

Power (kW)

Efficiency (%)

Measured

Design Adjusted

Design Original

90%

91%

92%

93%

94%

95%

96%

97%

98%

99%

100%

0 200 400 600 800 1000 1200 1400

Power (kW)

Efficiency (%)
77

5.3.2 Calorimetric Method

The calorimetric efficiency was calculated by estimating the following losses.

• Heat flux of the drivetrain case

• Conduction through the drivetrain stand

• Conduction through the gearbox lubrication system and generator coolant system

The heat flux of the drivetrain case was calculated using measurements taken from a thin-film 

heat-flux sensor, thermal imaging, and point temperature measurements. A heat-flux sensor was 

attached to the top of the drivetrain case, near the connection point between the generator and 

gearbox. Thermal imaging was used to estimate the average temperature of the gearbox and the 

drivetrain. The average temperature of the gearbox and generator was used in conjunction with 

the point heat-flux measurement and ambient temperature of the DTF to estimate the heat flux of 

the gearbox and generator cases. Figure 96 is an example of a thermal image of the drivetrain.

Figure 96. Thermal image example

The conduction of heat through the drivetrain stand was estimated by using thermal imaging to

determine the temperature drop across the torque arms of the stand. The geometry of the arms 

was used in conjunction with the temperature measurements to estimate the heat flux in the 

stand. Calculations indicated that only a small fraction of the drivetrain’s loss was conducted 

through the stand.

DNV calculated the conduction of the gearbox lubrication system and generator coolant system 

using the fluid supply and return temperatures and flow rate. The difference in the supply and 

return gearbox oil temperatures was between approximately 0°C and 1°C. The accuracy of the 

gearbox oil temperature measurements was insufficient to determine the temperature drop in the 

gearbox-oil coolant system. Because the loss calculations produced erroneous results, the heat 

loss in the gearbox-lubrication system was ignored in the final calculations.

It is difficult to separate the gearbox and generator heat loss because of the thermal coupling 

present. Figure 97, Figure 98, and Figure 99 provide plots of the measured and theoretical 
78

efficiency of the gearbox, generator, and the entire drivetrain. The maximum power level for 

which the generator was able to operate at thermal equilibrium with stator and magnet 

temperature at less than the design values was 1.0 MW. Calorimetric efficiency calculations, 

however, were performed for a 1.2-MW test run. A 1.2-MW test run was performed and, as the 

drivetrain approached thermal equilibrium, the thermal design limits of the drivetrain were 

exceeded slightly. Due to the high stator temperatures, this test run was stopped before thermal 

equilibrium was reached. Prematurely stopping the 1.2-MW test run should’ve yielded greaterthan-actual calorimetric efficiency.

Figure 97. Gearbox calorimetric efficiency

Figure 98. Generator calorimetric efficiency

90%

91%

92%

93%

94%

95%

96%

97%

98%

99%

100%

0 200 400 600 800 1000 1200 1400 1600

Power (kW)

Efficiency (%)

Measured

Design Adjusted

Design Original

90%

91%

92%

93%

94%

95%

96%

97%

98%

99%

100%

0 200 400 600 800 1000 1200 1400 1600

Power (kW)

Efficiency (%)

Measured

Design
79

Figure 99. Combined gearbox and generator calorimetric efficiency

5.4 Vibration Levels

The vibration was measured in three orthogonal directions on the east and west side of the 

generator (see Figure 7 for accelerometer locations). Figure 100 and Figure 101 are 

representative frequency spectrum plots of the vibration of the drivetrain at the east and west 

accelerometers at one third rated power. These spectra are dominated by the even harmonics of 

the electrical frequency and by the gearbox meshing frequency fundamental and second

harmonic. The electrical frequency at one third rated power is 53 Hz and the gearbox meshing

frequency is 319 Hz. The gearbox meshing frequency fundamental and second harmonics are 

circled in red in the frequency plots. The spectra of the western accelerometer contain frequency 

content not present at the eastern accelerometer, especially in the Y and Z directions. During 

testing, the lubrication pump was mounted on the west side of the drivetrain; this could be 

responsible for the additional spectral content.

90%

91%

92%

93%

94%

95%

96%

97%

98%

99%

100%

0 200 400 600 800 1000 1200 1400 1600

Power (kW)

Efficiency (%)

Measured

Design Adjusted

Design Original
80

Figure 100. Velocity at one third rated power, east

0 100 200 300 400 500 600 700 800 900 1000

10-4

10-2

100

Frequency (Hz)

Vibration (g), X

0 100 200 300 400 500 600 700 800 900 1000

10-4

10-2

100

Frequency (Hz)

Vibration (g), Y

0 100 200 300 400 500 600 700 800 900 1000

10-4

10-2

100

Frequency (Hz)

Vibration (g), Z
81

Figure 101. Vibration at one third rated power, west

Figure 102 and Figure 103 are representative frequency spectrum plots of the drivetrain vibration 

at the east and west accelerometers at rated power. The spectra are dominated by the fundamental and second harmonic of the meshing frequency of the gearbox—458 Hz and 917 Hz,

respectively, at rated power. The electrical frequency harmonics present at one third rated power

are significantly attenuated at rated power.

0 100 200 300 400 500 600 700 800 900 1000

10-4

10-2

100

Frequency (Hz)

Vibration (g), X

0 100 200 300 400 500 600 700 800 900 1000

10-4

10-2

100

Frequency (Hz)

Vibration (g), Y

0 100 200 300 400 500 600 700 800 900 1000

10-4

10-2

100

Frequency (Hz)

Vibration (g), Z
82

Figure 102. Vibration at rated power, east

0 100 200 300 400 500 600 700 800 900 1000

10-4

10-2

100

Frequency (Hz)

Vibration (g), X

0 100 200 300 400 500 600 700 800 900 1000

10-4

10-2

100

Frequency (Hz)

Vibration (g), Y

0 100 200 300 400 500 600 700 800 900 1000

10-4

10-2

100

Frequency (Hz)

Vibration (g), Z
83

Figure 103. Vibration at rated power, west

Figure 104 and Figure 105 show plots of the gearbox meshing fundamental and second 

harmonic. There appears to be a resonance in the drivetrain at 800 kW in the Y and Z directions.

The meshing fundamental vibration otherwise is flat.

0 100 200 300 400 500 600 700 800 900 1000

10-4

10-2

100

Frequency (Hz)

Vibration (g), X

0 100 200 300 400 500 600 700 800 900 1000

10-4

10-2

100

Frequency (Hz)

Vibration (g), Y

0 100 200 300 400 500 600 700 800 900 1000

10-4

10-2

100

Frequency (Hz)

Vibration (g), Z
84

Figure 104. Magnitude of gearbox meshing frequency fundamental

400 600 800 1000 1200 1400 1600

0

0.2

0.4

0.6

0.8

1

Power (kW)

X Vibration

(g)

400 600 800 1000 1200 1400 1600

0

0.2

0.4

0.6

0.8

1

Power (kW)

Y Vibration

(g)

400 600 800 1000 1200 1400 1600

0

0.2

0.4

0.6

0.8

1

Power (kW)

Z Vibration

(g)

East

West
85

Figure 105. Magnitude of gearbox meshing frequency second harmonic

400 600 800 1000 1200 1400 1600

0

0.2

0.4

0.6

0.8

1

Power (kW)

X Vibration

(g)

400 600 800 1000 1200 1400 1600

0

0.2

0.4

0.6

0.8

1

Power (kW)

Y Vibration

(g)

400 600 800 1000 1200 1400 1600

0

0.2

0.4

0.6

0.8

1

Power (kW)

Z Vibration

(g)

East

West
86

6 Conclusions

In general, the gearbox functioned as expected. The testing did not reveal any significant issues 

with the gearbox design or manufacture. The generator was able to operate at rated power but 

only for a short period before overheating occurred. The greatest power level for which the 

generator was able to achieve thermal equilibrium at less than the design limits was 1.0 MW.

During testing, the current and voltage levels of the generator were not consistent with the design 

specifications. The causes of the divergence between the measured and design values of current 

and voltage identified in this report are the lagging power factor of the generator, reduction in 

magnet strength with temperature, and increased voltage drop in the stator circuit. The power 

factor of the generator is immune to changes in temperature; however, the increased current 

caused by the lagging power factor caused some incremental increase in the loss in the windings. 

The interaction of the magnet temperature and the voltage and current levels formed a positivefeedback system that exacerbated electrical and thermal deviations from the design 

specifications.

Testing showed that the reworked series winding temperatures were significantly greater than the 

original parallel winding temperatures, even though the same cooling system was used for both 

winding configurations. The elevated series windings temperatures are the result of increased 

electrical resistance and thermal resistance. The increased electrical resistance amplifies the 

power loss in the windings, and increased thermal resistance reduces the ability of the cooling 

system to pull heat from the windings.

Calorimetric efficiency calculations showed greater than expected efficiency losses in the 

drivetrain, which suggests that the generator’s overheating is at least in part caused by increased 

efficiency loss. A portion of the increased loss can be attributed to the greater than expected 

current levels. The design of the series windings should be reviewed and compared with test 

data, however, to ascertain the cause of the generator overheating at power levels greater than 

1.0 MW.

The compact design of the drivetrain makes it more power dense than more conventional wind 

turbine drivetrain designs. The relatively high power density requires careful attention to thermal 

management. One of the lessons learned is that the design analysis should include perturbations 

on the system parameters to determine the sensitivity of the system to seemingly slight changes, 

especially in the case of positive-feedback systems.
87

Appendix A. Signal Lists

Table A-1. Instrumentation Connected to the Drivetrain Supervisory Controller and

Ported to the Dyno Data-Acquisition System (1 Hz)

Measurement Tag

Bearing lube temperature BLT

Main bearing temperature MBT

Dynamometer Test Facility ambient temp AT

Gearbox case temp GCT

Generator case heat flux HF

Generator coolant supply temp CST

Generator rotor coolant flow rate RFR

Generator rotor coolant return temp RRT

Generator stator coolant flow rate SFR

Generator stator coolant return temp SRT

Lube flow rate LFR

Lube pump pressure LPP

Lube return temperature LRT

Lube supply pressure LSP

Lube supply temperature LST

Main shaft position MSP

Main shaft speed MSS

Outside of coil end-turn segment #2 temp OET2

Inside of coil end-turn segment #2 temp IET2

End of laminate tooth segment #2 temp LTT2

Back iron laminate segment #2 temp BIT2

Cold plate segment #2 temp CPT2

Outside of coil end-turn segment #5 temp OET5

Inside of coil end-turn segment #5 temp IET5

End of laminate tooth segment #5 temp LTT5

Back iron laminate segment #5 temp BIT5

Cold plate segment #5 temp CPT5

Outside of coil end-turn segment #8 temp OET8

Inside of coil end-turn segment #8 temp IET8

End of laminate tooth segment #8 temp LTT8

Back iron laminate segment #8 temp BIT8

Cold plate segment #8 temp CPT8

Outside of coil end-turn segment #11 temp OET11

Inside of coil end-turn segment #11 temp IET11

End of laminate tooth segment #11 temp LTT11

Back iron laminate segment #11 temp BIT11
88

Measurement Tag

Cold plate segment #11 temp CPT11

Generator segment #2 temp SEG2

Generator segment #3 temp SEG3

Generator segment #5 temp SEG5

Generator segment #6 temp SEG6

Generator segment #7 temp SEG7

Generator segment #8 temp SEG8

Generator segment winding #9 temp SEG9

Generator segment winding #10 temp SEG10

Generator segment winding #11 temp SEG11

Rectifier heatsink #1 temp RHS1

Rectifier heatsink #2 temp RHS2

Table A-2. Instrumentation Connected to the FastDAS System (2 kHz)

Measurement Tag

Generator shaft torque GST

Generator shaft position GSP

Generator shaft index GSI

Inverter #1 real power IP1

Inverter #2 real power IP2

Main shaft torque MST

Main shaft speed MSS

Planet bearing load; 0 degrees upwind PBL1

Planet bearing load; 0 degrees center upwind PBL2

Planet bearing load: 0 degrees center PBL3

Planet bearing load: 0 degrees center downwind PBL4

Planet bearing load: 0 degrees downwind PBL5

Planet bearing load: 72 degrees upwind PBL6

Planet bearing load: 72 degrees center PBL7

Planet bearing load: 72 degrees downwind PBL8

Planet bearing roller load MRL

Planet bearing temperature PBT

Planet gear displacement inboard PGD3

Planet gear displacement lagging PGD4

Planet gear displacement leading PGD2

Planet gear displacement outboard PGD1

Planet roller load trigger MRT

Ring gear strain: 0 degrees upwind RGS1

Ring gear strain: 0 degrees downwind RGS2

Ring gear strain: 120 degrees upwind RGS3
89

Measurement Tag

Ring gear strain: 120 degrees downwind RGS4

Ring gear strain: 240 degrees upwind RGS5

Ring gear strain: 240 degrees downwind RGS6

Sun gear axial displacement SGA

Torque tube axial displacement TTA

Torque tube radial displacement TTR

Quadrant #1 phase A current Q1IA

Quadrant #1 phase B current Q1IB

Quadrant #1 phase C current Q1IC

Quadrant #1 line AB voltage Q1VAB

Quadrant #1 line BC voltage Q1VBC

Quadrant #1 line CA voltage Q1VCA

Rectifier #1 DC voltage R1VDC

Quadrant #2 phase A current Q2IA

Quadrant #2 phase B current Q2IB

Quadrant #2 phase C current Q2IC

Quadrant #2 line AB voltage Q2VAB

Quadrant #2 line BC voltage Q2VBC

Quadrant #2 line CA voltage Q2VCA

Rectifier #2 DC voltage R2VDC

Quadrant #3 phase A current Q3IA

Quadrant #3 phase B current Q3IB

Quadrant #3 phase C current Q3IC

Quadrant #3 line AB voltage Q3VAB

Quadrant #3 line BC voltage Q3VBC

Quadrant #3 line CA voltage Q3VCA

Rectifier #3 DC voltage R3VDC

Quadrant #4 phase A current Q4IA

Quadrant #4 phase B current Q4IB

Quadrant #4 phase C current Q4IC

Quadrant #4 line AB voltage Q4VAB

Quadrant #4 line BC voltage Q4VBC

Quadrant #4 line CA voltage Q4VCA

Rectifier #4 DC voltage R4VDC
90

Table A-3. Accelerometer Signals

Measurement Tag

Sensor 1 X direction X1

Sensor 1 Y direction Y1

Sensor 1 Z direction Z1

Sensor 2 X direction X2

Sensor 2 Y direction Y2

Sensor 2 Z direction Z2

Table A-4. Calculated Channels

Measurement Tag

Generator shaft power GSP

Generator shaft speed GSS

Main shaft power MSP

Quadrant #1 average RMS* current Q1I

Quadrant #1 average RMS voltage Q1V

Rectifier #1 RMS voltage R1V

Quadrant #1 RMS real power Q1P

Quadrant #1 RMS reactive power Q1Q

Quadrant #2 average RMS current Q2I

Quadrant #2 average RMS voltage Q2V

Rectifier #2 RMS voltage R2V

Quadrant #2 RMS real power Q2P

Quadrant #2 RMS reactive power Q2Q

Quadrant #3 average RMS current Q3I

Quadrant #3 average RMS voltage Q3V

Rectifier #3 RMS voltage R3V

Quadrant #3 RMS real power Q3P

Quadrant #3 RMS reactive power Q3Q

Quadrant #4 average RMS current Q4I

Quadrant #4 average RMS voltage Q4V

Rectifier #4 RMS voltage R4V

Quadrant #4 RMS real power Q4P

Quadrant #4 RMS reactive power Q4Q

*RMS = root mean squared