WIND TURBINE GEAR OILS PERFORMANCE EVALUATION€¦ · Test procedure The rolling bearings tests were performed using an oil volume of 14 ml (to promote a oil level up to the middle

Proceedings of the 6th International Conference on Mechanics and Materials in Design,

Editors: J.F. Silva Gomes & S.A. Meguid, P.Delgada/Azores, 26-30 July 2015

-1975-

PAPER REF: 5748

WIND TURBINE GEAR OILS PERFORMANCE EVALUATION

Ramiro Martins1(*), Carlos M.C.G. Fernandes

1, Jorge H. O. Seabra

2

1INEGI, University of Porto, Porto, Portugal 3Faculdade de Engenharia, University of Porto, Portugal (*)

Email: [email protected]

ABSTRACT

This work evaluates the performance of different ISO VG 320 wind turbine gear oils with

base oil formulations based in PAO, Mineral, Ester and PAG having quite different viscosity

indexes. The oils evaluation was done in rolling bearing tests and gearbox tests. Their

behaviour was compared for operating conditions in the range of the observable in a wind

turbine gearbox.

The experimental results showed considerable differences between the different oils and it

was observed that depending on the contact type the relative behaviour of the lubricants

would change, i.e. the best lubricant for the rolling bearing would not imply the best result on

the gearbox tests.

The gear geometry is also a very important parameter on a gear transmission torque loss. It

was observed that a decrease of up to 25% in torque loss could be achieved with a low loss

gear geometry.

Numerical models were implemented and calibrated with the purpose of obtain the friction

coefficient promoted by the tested oils and performing a better simulation of the mechanical

system.

Keywords: Gears, Rolling bearings, Friction coefficient, Lubricants, Efficiency.

INTRODUCTION

The reduction of power consumption and the improvement of equipment’s life is a constant

target for industry. A large percentage of the wind turbine installed use planetary gearboxes

and inside those gearboxes the two main type of components that dissipate energy are the

gears and the rolling bearing. Gears and rolling bearings power loss behavior are the main

focus on this work.

The most common torque loss sources occurring within the gearbox are: friction loss between

the meshing teeth (Höhn et al., 1996, Martins et al., 2005, Martins et al., 2006, Martins et al.,

2008, Martins et al., 2009, Magalhães et al., 2010, Marques et al., 2013, Fernandes et al.,

2014), friction loss in the rolling bearings (Eschmann et al., 1985, Hohn et al., 1996, 2005,

Fernandes et al., 2013, Fernandes et al., 2013), load independent gear losses (Ariura et al.,

1975, Mauz, 1987, Maurer, 1994, Höhn et al., 1996, Changenet et al., 2007, Changenet et al.,

2011, Marques et al., 2013) and energy loss due to air-drag (Diab et al., 2004, Seetharaman et

al., 2009). The seals also contribute for the power losses

The losses inside a gearbox have several side effects besides the immediate of decreasing

power output and one of the most important is the increment of the operating temperature of

the oil. The increase of oil temperature contributes to the increase of surface failure

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Tribology Trends for Higher Efficiency and Reliability

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probability, since wear, scuffing, micropitting or pitting (Hohn et al., 2004), due to thinner

film thickness and possible higher friction coefficient, nevertheless, the additive package

always plays a very important role.

ISO VG 320 is the European oil viscosity specification for wind turbine gearboxes. The ISO

VG 320 base oils have different base oils, different viscosity indexes, besides the typical oil

operating temperature is between 60 to 80ºC, but nevertheless the viscosity specification is for

40ºC. Another focus of this work is the analysis of the influence of different fully formulated

ISO VG 320 oils, taking into account their different viscosity, on the power losses of rolling

bearings and gears.

Five oils were selected, having all different base products.

LUBRICANTS

Five ISO VG 320 wind turbine lubricants were selected having different base oils. The

Lubricants have the following designations: MINR (mineral base), PAOR (polyalphaolefin

base), ESTR (ester base), PAGD (poly-alkylene-glycol) and MINE (poly-alkyl-methacrylate).

The lubricants physical and chemical properties were evaluated and are displayed in Table 1.

All the lubricants are fully formulated and rated for lubrication of wind turbine gearboxes.

Table 1 - Lubricants pgysical and chemical properties.

Parameter Unit MINR ESTR PAOR MINE PAGD

Base Oil [-]

Mineral Ester PAO Mineral

+ PAMA PAG

Chemical composition

Zinc (Zn) [ppm] 0.9 6.6 3.5 < 1 1

Magnesium (Mg) [ppm] 0.9 1.3 0.5 < 1 1.4

Posphorus (P) [ppm] 354.3 226.2 415.9 460 1100

Calcium (Ca) [ppm] 2.5 14.4 0.5 2 0.8

Boron (B) [ppm] 22.3 1.7 28.4 36 1.0

Sulphur (S) [ppm] 11200 406 5020 6750 362

Physical properties

Density @ 15 °C [g/cm3] 0.902 0.915 0.859 0.893 1.059

Viscosity @ 40 °C [cSt] 319.2 302.9 313.5 328.3 290.3

Viscosity @ 80 °C [cSt] 43.9 56.1 60.4 66.5 78.9

Viscosity @ 100 °C [cSt] 22.3 34.9 33.3 37.1 51.1

ROLLING BEARINGS TESTS

Two types of bearing were tested in this work, a 51107 Thrust Ball Bearings (TBB) and an

81107 Roller Thrust Bearings (RTB). These two bearing types were selected due to

promoting different contact pattern, elliptical contact on the TBB and linear contact on the

RTB.

The rolling bearings were tested in a test rig that allow to apply different operating conditions,

including controlled temperature and measure directly the rolling bearing torque loss.

The rolling bearing test assembly is displayed in Fi and a detailed description was published

by (Cousseau et al., 2011).



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Fig. 1 - Schematic view of the rolling bearing assembly for the Four-ball machine.

Test procedure

The rolling bearings tests were performed using an oil volume of 14 ml (to promote a oil level

up to the middle of the rolling elements) at a constant oil bath temperature of 80 ºC. The

operating conditions are detailed in Table 2.

Table 2 - Operating conditons in rolling bearings tests.

Speed [rpm] 75 till 1200

Oil temperature [ºC] 80

Load [N] 700 7000 700 7000

Load ratio (C/P) [/] 28.4 2.84 42 4.2

p0 [GPa] 1.15 2.48 0.32 1.0

Experimental results

Figures 2, 3, 4 and 5 represent the total torque loss measurement as function of the input

speed, for the two load levels tested and for the TBB and RTB rolling bearings tested.

For the TBB the torque loss increase with the increasing speed and the increase of load also

leaded to an increase of the torque loss, as expected. The increase of ten times on the load

only increased around three times the torque loss.

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Fig. 2 - Torque loss of TBB for a load of 700 N

The RTB total torque loss measurements for a load of 7000N and a speed of 1200rpm had a

operating temperature above 80ºC (the control of temperature does not have cooling

capability).

The torque loss measured with TRB was much higher that the measured for TBB. For the

RTB it is observed an increase of torque loss with increasing speed for low

conditions(Figure 4) while for high load conditions a decrease of torque loss is observed with

increasing speed (Figure 5).

Fig. 4 - Torque loss of RTB for a load of 700 N

Discussion

The TBB results show that the MINR and PAGD lubricants had the higher torque loss from

the group, while PAOR had the lowest torque loss for the load of 700N and MINE had the

lowest torque loss for a load of 7000N. The ESTR oil had an intermediate behaviour for all

tests.

For RTB tests there are significa

increases from 700 to 7000 N. For the load of 700N the PAGD lubricant displays the largest

toque loss while the MINR and ESTR had the lowest, result that appears to

lubricants viscosity. For the load of 7000 N the PAGD lubricant had the lowest torque loss for

the lower speeds while the mineral bases had the highest.

The relative behaviour of the lubricants is dependent on the bearing type, being the slid

speed the biggest difference between the TBB and the RTB, and also the contact pressure.


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Torque loss of TBB for a load of 700 N Fig. 3 - Torque loss of TBB for a load of 7000 N


erating temperature above 80ºC (the control of temperature does not have cooling


RTB it is observed an increase of torque loss with increasing speed for low

(Figure 4) while for high load conditions a decrease of torque loss is observed with

Torque loss of RTB for a load of 700 N Fig. 5 - Torque loss of RTB for a load of 700 N

sults show that the MINR and PAGD lubricants had the higher torque loss from



For RTB tests there are significant changes on the lubricants behaviour when the load


toque loss while the MINR and ESTR had the lowest, result that appears to


the lower speeds while the mineral bases had the highest.

The relative behaviour of the lubricants is dependent on the bearing type, being the slid


Torque loss of TBB for a load of 7000 N


erating temperature above 80ºC (the control of temperature does not have cooling


RTB it is observed an increase of torque loss with increasing speed for low load

(Figure 4) while for high load conditions a decrease of torque loss is observed with

Torque loss of RTB for a load of 700 N

sults show that the MINR and PAGD lubricants had the higher torque loss from



t changes on the lubricants behaviour when the load


toque loss while the MINR and ESTR had the lowest, result that appears to be driven by


The relative behaviour of the lubricants is dependent on the bearing type, being the sliding



Editors: J.F. Silva Gomes & S.A. Meguid, P.Delgada/Azores, 26

Figure 6 and Figure 7 display the specific film thickness calculated for the tests performed for

TBB and RTB, respectively. The first fact to be noticed is that T

under boundary lubrication regime, while for the RTB half of the tests were performed in

boundary and half in mixed lubrication regime.

Fig. 6 - Specific film thickness for TBB bearing with

an applied load of 7000 N

FZG GEAR TESTS

The well-known FZG test rig used

has a torque transducer installed (

gearboxes.

Fig. 8 - Schematic representation of the FZG test rig with the input torque (Tin) and torque loss (TL) measured by

A standard spur gear geometry w

geometrical characteristics are represented in

Table 3 - Geometrical properties of the spur and helical tested gears.

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display the specific film thickness calculated for the tests performed for

TBB and RTB, respectively. The first fact to be noticed is that TBB tests were performed


boundary and half in mixed lubrication regime.

Specific film thickness for TBB bearing with Fig. 7 - Specific film thickness for RTB bearing with

an applied load of 7000 N

used in this work (Winter et al., 1985). The FZG test rig used

installed (Fig), allowing to measure the total torque loss (

Schematic representation of the FZG test rig with the input torque (Tin) and torque loss (TL) measured by

the torque meter.

A standard spur gear geometry was tested and also two helical gear geometries. The gears

geometrical characteristics are represented in Table .

Geometrical properties of the spur and helical tested gears.

display the specific film thickness calculated for the tests performed for

BB tests were performed


Specific film thickness for RTB bearing with

. The FZG test rig used

allowing to measure the total torque loss (TL) on both

Schematic representation of the FZG test rig with the input torque (Tin) and torque loss (TL) measured by

as tested and also two helical gear geometries. The gears

Geometrical properties of the spur and helical tested gears.

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Test procedure

The drive gearbox (near the electric motor) has a 40 mm width spur gear and four NJ406

cylindrical roller bearings and it was the same during all tests.

The same type C40 spur gear set was used for testing all the lubricants. The test gearbox

shafts were supported with four NJ406 cylindrical roller bearings when spur gears are tested

(the drive gearbox always use this configuration).

For the test of the helical gear a different case gearbox is used due to the axial forces

generated by the helical gears. The test gearbox has two NJ406 cylindrical rolling bearings

and two QJ308 four point contact ball bearing.

The operating condition used for the gear torque loss tests are represented in Table 4,

including the axial forces as well as an approximate value of the maximum Hertzian contact

pressure for the spur and helical gears.

Four load stages were selected, one (K1) with a very small load to study the churning losses

and the other to study the influence of load at contact pressures similar to the observed in real

wind turbine gearboxes. Three speed were selected, promoting tangential speed on the contact

similar to the observed in each of the wind turbine gearbox multiplier stages.

The gearbox was lubricated using oil injection with a flow of 3 l/min at a constant temperature

of 80 ºC.

The test procedure can be summarize as:

1. Run load stage Ki at each input speed condition (Table 4) during 3h;

- Continuous registry of the assembly working temperatures;

- Continuous torque measurement;

2. Increase the load stage and repeat procedure till the highest load stage

Table 4 - Operating conditions inn the FZG torque loss tests on spur and helical gear tests.

Wheel Speed [rpm]

200 400 1200 Spur C40 Helical H501/H951

FZG

load

stage

Wheel

torque Input power Fr Fa pH Fr Fa pH

Ki [N.m] [W] [N] [N] [MPa

] [N] [N] [MPa]

K1 4.95 104 207 622 37 0 108 37 24 171

K5 104.97 2198 4397 13191 790 0 498 777 518 787

K7 198.68 4161 8322 24967 1495 0 685 1471 980 1083

K9 323.27 6771 13541 40623 2432 0 874 2393 1594 1382



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The results presented for torque loss measurements and temperature are the average of the last

30 minutes of operation, i.e. the steady state operating conditions.

Experimental results

Table 5 displays the measured torque loss for load stages K1 and K5. It is interesting to

observe that gear H501 displays the lowest torque loss at K1 and that H951gear has roughly

the same torque loss as C40 gear.

At load stage K5 the helical gears already display lower torque loss than the spur gear.

For the highest load stages (K7 and K9), the torque loss measurements are displayed in Table

6. At these load levels the helical gears display lower torque losses, especially the gear H951.

The synthetic lubricants also have highest decrease on torque loss with the helical geometry.

The MINE tests with H951 geometry couldn’t be finished due to a test rig failure.

Discussion

As the film thickness is an important parameter for the understanding of the torque loss and it

is a function of the viscosity and piezoviscosity of each lubricant. The tests were performed at

a constant temperature of 80 ºC. Table 1 displays the viscosity and piezoviscosity for each

lubricant at 80ºC. It is clear that the viscosity of the lubricants has huge differences, with the

PAGD exhibiting almost twice the viscosity of MINR.

This viscosity difference should be immediately reflected in the load stage K1torque loss, as

observed in C40 results, although those viscosity differences are not clear for helical gear

geometries.

The biggest differences observed between oil formulations are observed for the load stages

performed with highest loads (see Table 6). For those conditions, independently of the gear

geometry, the PAGD oil always had the lowest torque loss for the two lowest speeds while

the MINR displayed the highest torque loss for all the tests.

The MINE had in general torque loss results closer to the mineral base than PAGD, while

with PAOR and ESTR the torque loss results were close to the PAGD oil than the mineral

base.

The effect of speed on the torque loss is clearly observed with PAGD oil that increases the

torque loss at a faster pace than the other oils due to its higher viscosity and piezoviscosity.

The observation of the measured torque loss evolution with the increasing speed shows that

for MINR, PAOR and ESTR the minimum value is obtained for 400 rpm at load stage K7

while at load stage K9 the minimum is obtained somewhere between 400 rpm and 1200 rpm,

although for PAGD the minimum value was always obtained for input speeds lower than

400 rpm or even lower than 200 rpm. This behaviour clear indicated that the optimum

operating conditions for PAGD took place at smaller speed, i.e. it generates thicker lubricant

films for lower input speed. It is also clear that for high input speed the PAGD promotes very

high losses due to churning losses both in rolling bearings and gears.

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Table 5 - Evolution of the measured torque loss with speed for load stages K1 and K5. K1 K5

C40

H501

H951

Power loss numerical model

To extend the comprehension of the experimental results presented and to allow the forecast

of power losses in a transmission, a numerical model has been developed and implemented to

calculate the different forms of torque loss inside a gearbox for different operating conditions,

gear geometries and different lubricants (Magalhães et al., 2010, Martins et al., 2011, Martins

et al., 2011, Magalhães et al., 2012, Fernandes et al., 2013, Fernandes et al., 2013, Fernandes

et al., 2013, Fernandes et al., 2013, Fernandes et al., 2014, Marques et al., 2014).

Fig displays an equation presented by (Höhn et al., 2009) including all the sources of power

losses in a gear transmission. There are power losses dependent on load and also no-load

dependent power losses. This division is usually made for gears and bearings. The equations

mainly states that the total power losses (PV) are the sum of: gear no-load losses (PVZ0); load

dependent losses gear losses (PVZP); bearings losses (PVL); seal losses (PVD) and some eventual

auxiliary losses.

The main focus on this work was gears and bearings losses, being those characterized for each

lubricant, gear geometry and rolling bearing used in order to do an accurate fit of the



-1983-

experimental results gathered and to allow a forecasting of power losses for different

operating conditions and gear/bearing geometries.

Table 6 - Evolution of the measured torque loss with speed for load stages K7 and K9.

K7 K9

C40

H501

H951

Fig. 9 - Sources of power losses in a transmission

Fernades et al. (Fernandes et al., 2013, Fernandes et al., 2013, Fernandes et al., 2014) detailed

the numerical model and the calculation procedure, so only some remarks will be presented in

this work.

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Bearing power losses

The bearings power losses were calculated using the SKF numerical model (2005). The

experimental results presented for the RTB and TBB rolling bearings allowed the

determination of the two coefficients of friction parameters (µbl and µEHL) to be used in the

numerical model. Those parameters are lubricant dependent and must be determined for linear

and elliptical contacts i.e. roller thrust bearing (RTB) and thrust ball bearings (TBB),

respectively. Table displays the value of each coefficient of friction for each oil tested,

obtained with the experimental results displayed. These coefficient obtained for each oil

tested, allow the determination of the bearings power losses for different bearing sizes,

different geometries at different operating conditions with much better accuracy.

Table 5 - Boundary (µbl) and EHD (µEHL) coefficient of friction for TBB and RTB bearings at 80ºC.

Gears power losses

With the numerical model of the rolling bearings calibrated, it was possible to use the

experimental results obtained with the FZG gear tests to obtain the friction coefficient

between gear teeth minimizing the difference between the numerical power loss calculation

and the experimental results.

The numerical determination of the gears power losses were performed according to Hohn et

al. (Höhn et al., 1996). On this formulation special attention must be paid to the gear loss

factor (HV) once this is the factor that distinguishes the different gear geometries. The load

dependent power losses on the gears (PVZP) are calculated according to (Equation 1). The gear

loss factor is calculated according to (Equation 2) and is also dependent on the load

distribution along the path of contact, especially for the helical gears. The average friction

coefficient between teeth contact is determined according to (Equation 3) that is dependent on

the lubricant parameter (XL).

The lubricant parameter (XL) is the parameter that translates the influence of the lubricant on

the gear friction. The (XL) was obtained through the experimental tests presented.


Editors: J.F. Silva Gomes & S.A. Meguid, P.Delgada/Azores, 26

Table 6 displays the lubricant parameter cal

operating conditions tested. With this parameter it is possible to calculate the average friction

coefficient for different operating conditions.

Table 6 -

Numerical model results

The comparison of the experimental power loss measurements and the numerical model are

displayed in Figure 10for the MINR lubricant, showing an excellent correlation (>99.5%).

The numerical model allows to distinguish between gear

different operating conditions.

Fig. 10 - Comparison of the experimental power losses with the numerical results obtained with the numerical

al Conference on Mechanics and Materials in Design,


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displays the lubricant parameter calculated for the oils tested, valid within the


coefficient for different operating conditions.

- Lubricant parameter determined for the oils tested.


for the MINR lubricant, showing an excellent correlation (>99.5%).

The numerical model allows to distinguish between gear geometries and lubricants for

different operating conditions.

Comparison of the experimental power losses with the numerical results obtained with the numerical

model for the MINR lubricant.

(1)

(2)

(3)

culated for the oils tested, valid within the



for the MINR lubricant, showing an excellent correlation (>99.5%).

geometries and lubricants for

Comparison of the experimental power losses with the numerical results obtained with the numerical

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CONCLUSION

The experimental results allowed to observe the influence of different oil formulations on

rolling bearings and gears power losses, it also allowed to infer the influence of an optimized

gear geometry on the power losses.

The combination of the best gear geometry with the best oil formulation the global efficiency

of the gearbox was improved over 1%.

The power loss model, using the inputs determined using the RTB and TBB torque loss tests

and the FZG power loss tests, proved to accurately predict the experimental results. The use

of the two rolling bearings friction coefficients (µbl and µEHL) and the XL lubricant parameter

are fundamental for a correct prediction of the numerical model.

ACKNOWLEDGMENTS

The authors gratefully acknowledge the funding by - Fundação para a Ciência e a Tecnologia

within the project EXCL/EMS-PRO/0103/2012. This work was co-funded by: COMPETE,

QREN, EU.

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Documents

WIND TURBINE GEAR OILS PERFORMANCE EVALUATION€¦ · Test procedure The rolling bearings tests were performed using an oil volume of 14 ml (to promote a oil level up to the middle