Torque loss and wear of FZG gears lubricated with wind turbine gear oils using an ionic liquid as additive

Author's Accepted Manuscript

Torque loss and wear of FZG gears lubricated with wind turbine gear oils using an ionic liquid as additive Carlos M.C.G. Fernandes, Antolin Hernandez Battez, Rubén González, Raquel Monge, José L. Viesca, Alberto García, Ramiro C. Martins, Jorge H.O. Seabra www.elsevier.com/locate/triboint

PII: DOI: Reference:

S0301-679X(15)00189-9 http://dx.doi.org/10.1016/j.triboint.2015.04.037 JTRI3665

To appear in:

Tribology International

Received date: 27 February 2015 Revised date: 9 April 2015 Accepted date: 17 April 2015 Cite this article as: Carlos M.C.G. Fernandes, Antolin Hernandez Battez, Rubén González, Raquel Monge, José L. Viesca, Alberto García, Ramiro C. Martins, Jorge H.O. Seabra, Torque loss and wear of FZG gears lubricated with wind turbine gear oils using an ionic liquid as additive, Tribology International, http: //dx.doi.org/10.1016/j.triboint.2015.04.037 This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting galley proof before it is published in its final citable form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.

Torque loss and wear of FZG gears lubricated with wind turbine gear oils using an ionic liquid as additive Carlos M. C. G. Fernandesa,1,∗, Antolin Hernandez Battezb,c , Rub´en Gonz´alezd,c , Raquel Mongee , Jos´e L. Viescab,c , Alberto Garc´ıab , Ramiro C. Martinsa , Jorge H. O. Seabraf a INEGI,

Universidade do Porto, Campus FEUP, Rua Dr. Roberto Frias 400, 4200-465 Porto, Portugal b Department of Construction and Manufacturing Engineering, University of Oviedo, Asturias, Spain c Faculty of Science and Technology, Bournemouth University, UK d Department of Marine Science and Technology, University of Oviedo, Asturias, Spain e Engineering Division, CELLS - ALBA Synchrotron Light Facility, Barcelona, Spain f FEUP, Universidade do Porto, Rua Dr. Roberto Frias s/n, 4200-465 Porto, Portugal

Abstract This work presents a study of the tribological behaviour of a mineral fullyformulated wind turbine gear oil additised with [BMP][NTf2 ] ionic liquid. The target application are the wind turbine gearboxes, thus the fully formulated oil with and without ionic liquid additive was tested in a rolling bearings test rig to measure the thrust rolling bearing torque loss and was also tested in a FZG gear test rig to measure the gears torque loss at operating conditions similar to the observed in a wind turbine gearbox. The results show that a wind turbine gear oil additised with ionic liquid can reduce the torque loss and improve the gearbox efficiency while producing less wear particles as observed in the oil analysis. Keywords: ionic liquids, additives, lubrication, wind turbine gear oils

1. Introduction Reduction of CO2 emissions is one of the most important concerns in today’s society due to the climate change. Wind energy has been playing an important ∗ Corresponding

author Email address: [email protected] (Carlos M. C. G. Fernandes)

Preprint submitted to Tribology International

April 24, 2015

role on the electricity generation, specially during the last years, with the emergence of new designs of wind turbines with increasing size and output power. For large wind turbines, the low rotor speed makes necessary the usage of multiple stage gearboxes, in order to multiply the speed for the adequate rotating speed of the generator [1]. The higher repair and maintenance costs of the wind turbines are concentrated in the gearbox. Furthermore, the downtime due to gearbox failure is usually greater than any other component failures. This aspect caused the emergence of new direct drive wind turbine designs, but without better reliability results than gearbox drive turbines [2]. In wind turbine gearboxes, one of the most important tasks is the selection of the lubricants, because the right viscosity and anti-scuffing properties of oils are reached at temperatures above 80 ◦ C. However the operation temperatures in wind turbines gearboxes are usually lower than 60 ◦ C, promoting lower antiscuffing protection and resulting in higher start up wear [3]. On the other hand, above 80 ◦ C is necessary to use synthetic oils, because it is very difficult to maintain an effective lubricant film [3]. Approximately 30% of the energy losses in wind turbine gearboxes are caused by the rolling bearings. However, at high operation torque the friction losses between the meshing teeth are the main source of power loss. These energy losses depend on the rheological and film forming properties of the gear oil [4, 5, 6, 7]. Fernandes et al. [8] presented the torque loss in a FZG gearbox using five commercial fully-formulated gear oils with ISO VG 320 grade: a mineral-based gear oil, an ester-based gear oil, a polyalphaolefin based gear oil, a mineral + poly(alkyl methacrylate) gear oil and a polyalkylene glycol based gear oil, obtaining substancial differences in the torque loss behaviour for each lubricant. One of the most important challenges of the industry consists in improve tribological behaviour of mechanical transmissions, and for this objective, the research of high performance gearbox oils using new additives is one of the most promising approaches. Ionic liquids have been proposed many times as oil 2

additives in the development of novel high performance lubricants [9]. Ionic liquids (ILs) are molten salts at room temperature, composed by organic or inorganic anions and cations. ILs are used in the lubrication of several mechanical contacts in vacuum applications [10] due to their main properties: low vapour pressure, thermal and chemical stability, high thermal and electrical conductivity and tunable properties. Ye et al. [11] published the first research work on the use of ILs on lubrication showing the good tribological properties of different imidazoliumbased ILs for various material contacts in pure sliding conditions. Different material pairs such as steel/steel, steel/SiO2 , steel/aluminium, steel / copper, steel/Si(100), steel/sialon, steel/titanium, sialon/Si3 N4 and steel-PVD coatings (TiN, DLC and CrN) have been explored using different cations (pyridinium, imidazolium, phosphonium and ammonium) combined with anions such as hexafluorophosphate (PF6 ), tetrafluoroborate (BF4 ), CF3 SO3 , (C2 F5 )3 PF3 (or FAP) and (CF3 SO2 )2 N (or NTf2 ) [12, 13, 14, 15, 16, 17, 18, 19, 20, 21, 22, 23, 24, 25, 23, 26, 27, 28, 29]. The most frequently anions studied in the tribology literature are PF6 and BF4 . However, their tendency to produce HF (very corrosive) and hydrolytic instability have been reported [29]. A new type of ILs using the FAP anion (manufactured by Merck KGaA [30]) shows a better hydrolytic stability. Several review papers [31, 32, 33, 34] have explained the advantages of using ILs in lubrication, their oxidative and thermal stability, tribochemical reactions and tribological properties. ILs can be used as neat lubricants, or as additive to base or formulated oils. The use of ILs as pure lubricant presents the problem of the high cost of ILs compared with any mineral or synthetic hydrocarbon oils. The use of ILs as additive in oil is a cheaper solution, but could present a problem of low solubility in non-polar oils [35]. The wear and friction behaviour of ILs as oil additive or as neat lubricant have been studied using different mineral and synthetic base oils [29, 35, 36, 37]. The use of ILs as additional additives of fully formulated oils needs to be 3

studied in depth. The combined effect of the IL with another existing oil additives has been studied by Qu et al. [38] reporting the improved performance of SAE 5W-30 engine oil using a phosphonium-based ionic liquid with concentration of 5 wt%. Greco et al. [39] obtained similar results using 1 wt% of nano-colloidal boron nitride as additive in fully formulated wind turbine gearbox oil. Monge et al. [40] have studied the friction and wear behaviour of two [NTf2 ] anion-based ionic liquids used as additive to two fully formulated (polyalphaolefinand mineral-based) wind turbine gear oils, using ball on plate reciprocating tests. Results indicated a slightly friction reduction; however the use 5 wt% of both ILs improved the wear performance of the gear oils under all test conditions employed. Based on the good anti friction behaviour, hydrophobicity and hydrolytic stability of the [NTf2 ] anion reported by Minami et al. [18, 32], the current work studies using a new approach the tribological performance of the best mixture found in [40], the [BMP][NTf2 ] ionic liquid (1-butyl-1-methylpyrrolidinium bis (trifluoromethylsulfonyl) imide) used as additive to a mineral-based fully formulated wind turbine gear oil. The study is based on the power loss of FZG gear tests and rolling bearing tests. Additionally to the efficiency improvement, the wear protection provided by the addition of the ionic liquid is also studied.

2. Lubricants A commercial ISO VG 320 fully formulated mineral-based gear oil (MINR) was tested in this work. This oil is widely used in wind turbine gearboxes and in a previous work [40] it was tested with and without ionic liquids in its composition using a different tribological approach. The chemical composition and rheological properties was previously determined by Fernandes et al. [4, 5, 6, 7, 8] and are displayed in Table 1. 1-butyl-1-methylpyrrolidinium bis (trifluoromethylsulfonyl) imide, [BMP][NTf2 ], provided by (Io-Li-Tec), was used as additive in this paper. The main properties

4

Table 1: Physical and chemical properties of wind turbine gear oil used.

Parameter

Unit

MINR

[-]

Mineral

Zinc (Zn)

[ppm]

0.9

Magnesium (Mg)

[ppm]

0.9

Phosphorus (P)

[ppm]

354.3

Calcium (Ca)

[ppm]

2.5

Boron (B)

[ppm]

22.3

Sulphur (S)

[ppm]

11200

[g/cm3 ]

0.902

ν @ 40 C

[cSt]

319.22

ν @ 70◦ C

[cSt]

65.81

[cSt]

22.33

[-]

85

Base Oil Chemical composition

Physical properties ρ @ 15◦ C ◦

◦

ν @ 100 C VI

of this ionic liquid are shown in Table 2, including the molecular structure. Before each experimental test, ionic liquid at 5 wt% was blended with the mineral oil, using a mechanic mixer for 10 minutes and controlling the temperature of the samples at 70 ◦ C. The corrosion activity, thermogravimetric performance and rheological properties of the mixture was previosuly studied in [40]. On the one hand, the surface analysis by optical microscopy, SEM and EDS after testing did not show corrosion activity. On the other hand, the thermogravimetric analysis (TGA) made to the blend and the gear oil without ionic liquid showed that thermal stability (decomposition temperature of around 230 ◦ C) was not improved with the addition of the ionic liquid. Finally, the addition of the ionic liquid resulted in a small increase in viscosity and viscosity index of the blend with regard to the gear oil without ionic liquid. All these results meet some of requirements 5

Table 2: Main properties and molecular structure of the ionic liquid. Ionic Liquid Cation

[BMP]

Anion

[NTf2 ]

IUPAC name

1-Butyl-1-

Purity

Density

Viscosity*

Viscosity

(%)

(g/cm3)

(mm2 /s)

Index*

> 99

15 ◦ C

40◦ C

100 ◦ C

1.40

28.826

6.228

174

methylpyrrolidinium bis(

trifluoromethyl-

sulfonyl) imide *Measured in a SVM 3000 Stabinger viscometer (ASTM D7042, D2270)

Molecular structure C9 H20 N

(CF3 SO2 )2 N−

[BMP]

[NTf2 ]

established by DIN 51517 Part 3; however, for a commercial use, further analysis should be done in order to verify all the standard requirements and possible interactions of the ionic liquid with other lubricant additives.

3. Rolling bearings torque loss 3.1. Test rig The tests were performed on a special test rig designed to perform friction torque measurements in rolling bearings, as shown in Figure 1. The rolling bearing assembly is divided in two parts: the shaft adapter (6), directly connected to the machine shaft and supporting the bearing upper race (5); a lower race support (2) and the bearing lower race (3), both clamped to the bearing housing (1). In operation, the internal bearing torque (or friction torque) is transmitted to the torque cell (11) through the bearing housing (1). The friction torque was measured with a piezoelectric torque cell KIST LER

6

9339A, ensuring high-accuracy measurements even when the friction torque generated in the bearing was very small compared to the measurement range available. This assembly has five thermocouples (I-V), measuring temperatures at strategic locations (see Figure 1) which are used to monitor the temperature inside the bearing assembly (IV), near to the rolling bearing and the lubricant (III) and to evaluate the heat evacuation from the bearing housing into the surrounding environment (I, II and V). The system is also monitored by two thermocouples to quantify the chamber and room temperatures. When assembled in the modified four-ball machine the rolling bearing assembly is submitted to a continuous air flow, forced by two fans with 38 mm in diameter and running at 2000 rpm, cooling the chamber surrounding the bearing house. The rolling bearing is lubricated by an oil volume of 14 ml. This volume was selected so that the oil level reaches the centre of the roller, such as advised by the manufacturer [41]. The tests were performed with a cylindrical roller thrust bearing (RTB 81107) that is suitable for arrangements that have to support heavy axial loads. A linear contact is generated between the cylindrical roller and the rings. For each oil tested a new rolling bearing was used in order to avoid possible interactions between oil samples and also try to keep similar surface roughness. 3.2. Test procedure All thrust bearing tests were performed applying two different axial loads of 700 N and 7000 N at five different rotational speeds: 150, 300, 600, 900 and 1200 rpm. The operating conditions are presented in Table 3 for the RTB 81107 rolling bearing. The machine was started at the desirable speed and run until it reached a constant temperature. Under stabilized conditions, four friction torque measurements were performed: three values are stored and the most dispersed one was disregarded. Due to the “drift effect”, which affects the measurements of 7

8

I

7

1 - Bearing house; 2 - Lower race support; 3 - Bearing lower race; 4 - Rolling element and cage assembly; 5 - Bearing upper race; 6 - Shaft adapter; 7 - Retainer; 8 - Cover; 9 - Upper protecting plates; 10 - Upper connection pins; 11 - Torque cell; 12 - Lower connection pins; 13 - Lower protecting plates;

n

6 B 5

UP

4 C 3 A

2 1

I - Cover temperature; II - Bearing house temperature; III - Oil temperature; IV - Internal temperature; V - Inferior temperature;

II

LW

9

III

10

IV

11

P - Load n - Rotational speed

V

12 13 P

Figure 1: Schematic view of the rolling bearing assembly.

the piezoelectric sensors after long periods of operation, the friction torque measurements should be made in a short period of time (less than 120 s) and at constant temperature (± 2 ◦ C). 3.3. Total torque results The total torque loss measurements obtained in the tests performed with 700 N and a cylindrical roller thrust bearing RTB are presented in Figure 2a. Table 3: Roller-raceway contact parameters (RTB 81107).

Axial Load [N ]

700

RX [m]

5.00 × 10−3

L [mm]

5.00

ah [μm] p0 [M P a]

8

7000

13.76

43.50

318

1004

MINR

150

MINR+5% IL

125

MINR+5% IL

450 Mt [Nmm]

100

Mt [Nmm]

MINR

500

75

50

350 7000 N

300

700 N

25

400

250

0 0

200

400 600 800 1000 Rotational speed [rpm]

0

1200

200

(a) 700 N.

400 600 800 1000 Rotational speed [rpm]

1200

(b) 7000 N.

Figure 2: Total torque loss TL vs. rotational speed for RTB 81107 rolling bearing at 80 ◦ C. Table 4: Oil temperature at 1200 rpm and 7000 N.

θoil [◦ C]

Δθ [◦ C]

MINR

93.3

62.0

MINR+5% IL

91.9

58.9

Oil

The addition of 5 % of ionic liquids to the mineral oil (MINR+5% IL), under an axial load of 700 N, promoted a small reduction in the torque loss. The total torque loss measurements under an axial load of 7000 N, are presented in Figure 2b. In the case of the MINR with ionic liquids (MINR+5% IL) at rotational speeds higher than 900 rpm the friction torque loss is considerably lower than the obtained for the mineral oil (MINR). The oil temperature at 900 and 1200 rpm, under an axial load of 7000 N, was higher than 80 ◦ C for both lubricants. Table 4 presents the oil operating temperature and stabilization temperature for 1200 rpm and 7000 N. The stabilization temperature is given by equation (1). The gear oil (MINR) additised with ionic liquids reached a stabilization temperature lower than MINR without ionic liquids. The reduction in the stabilized temperature is close to 3 ◦ C. Δθ = θoil − θchamber

9

(1)

(a) Schematic view.

(b) Test gearbox and(c) Torque transtest gear.

ducer.

Figure 3: FZG gear test rig.

4. Torque loss on FZG gears 4.1. Test rig The FZG test machine used in this work is presented in the Figure 3. The FZG machine is a gear test rig operating with a static torque in a power recirculating concept [42]. The drive gearbox (3) connects the test pinion (1) and wheel (2) through two shafts. The test pinion shaft is separated into two parts by the load clutch (4). One part of the clutch can be locked using the locking pin (5), while in the other part of the clutch different static torques can be applied using the load lever and dead weights (6). The torque loss (TL) was measured using a torque transducer (ETH Messtechnik DRDL II) integrated in the FZG test rig, as shown in Figure 3c. A sensor interface (ValuemasterBase) is used by the system to communicate with a computer using an Ethernet port. The software and hardware of torque cell allows measuring and recording the torque with a sampling rate between 1 Hz and 1000 Hz. The temperatures of the test were measured using eight different thermocouples located in different points of the machine. Type C gear with a face width of 40 mm was assembled in the FZG drive 10

Table 5: Geometry of the FZG gears tested.

Gear type

Number of teeth

type C40

type C14

Pinion

Wheel

Pinion

Wheel

16

24

16

24

Module [mm]

4.5

Centre distance [mm]

91.5

Pressure angle [◦ ]

20

Working pressure angle [◦ ]

22.44

Helix angle [◦ ]

0

Contact path [mm]

19.0987

Face width [mm]

40

14

Addendum modification [/]

+0.1817

+0.1715

+0.1817

+0.1715

Addendum diameter [mm]

82.64

118.54

82.64

118.54

Transverse contact ratio α [/]

1.44

Overlap contact ratio β [/]

0

Material

20MnCr5

Ra [μm]

0.7

≈ 0.5

gearbox while a type C gear with face width of 14 mm was assembled in the test gearbox. Table 5 displays the main dimensions of both gears. 4.2. Test conditions and test procedure Three input speeds and four load stages were preformed in FZG test campaign. The operating conditions used in the torque loss tests are displayed in Table 6. The tangential speed, the power circulating in the system, the tangential force transmitted by the gears, the radial forces on the rolling bearings and the Hertzian pressure in the gears are also included. The slave gearbox was oil jet lubricated with an oil flow of 3 l/min at 80 ◦

C always with MINR oil without ionic liquids. The test gearbox was dip

lubricated with 1 l of the candidate oils (MINR or MINR+5% IL) at 80 ◦ C. 11

Table 6: Operating conditions on the torque loss tests for C14 test gears and C40 driven gears. load stages with a load lever arm of 0.35 m Load stage

K1

Torque

Wheel Speed

vt

Power

Fbn

Fr

p0 C14

p0 C40

(N m)

(rpm)

(m/s)

(W )

(N )

(N )

(M P a)

(M P a)

200

1.1

103.7

400

2.3

207.3

98

49.5

188

111

1200

6.8

622.0

200

1.1

2198.5 2069

1049

511

865

3915

1986

704

1189

6371

3231

898

1517

4.95

K5

104.97

K7

198.68

K9

323.27

400

2.3

4396.9

1200

6.8

13190.9

200

1.1

4161.2

400

2.3

8322.4

1200

6.8

24967.2

200

1.1

6770.4

400

2.3

13540.9

1200

6.8

40622.7

The test procedure can be summarized as follows: 1. Run load stage Ki at each rotational speed Ni condition (Table 6) during 4 h according to the test sequence presented in Figure 4 and: • Register the assembly working temperatures. • Continuous torque measurement with a sample rate of 1 Hz. 2. Repeat procedure till the highest load stage (K9). The values reported for torque loss and temperature are the average of the last 30 min of operation (only the steady state operating conditions are considered for the average calculation). Between each oil test, the gearboxes were flushed with solvent. 4.3. Torque loss results The total torque loss measurements for all the test conditions performed with C40 gears assembled on drive gearbox and C14 on test gearbox are presented in Table 7. 12

200 rpm (4 hours)

Start - Ki

400 rpm (4 hours)

End

1200 rpm (4 hours)

Next Load Stage

No

Yes

Load stage K9?

Figure 4: Test sequence.

It is important to mention that the slave gearbox was always lubricated under oil jet lubrication with MINR. The total torque loss was previously measured with C40 gears installed in both the test and drive gearboxes and presented in Table 8 (procedure presented in [8]). The torque loss of the test gearbox (TLC14 ) is then given by Equation (2). TLC14 = TLC40+C14 −

TLC40+C40 2

(2)

Figure 5a displays the torque loss measured for load stage K1 at the input speeds of 200, 400 and 1200 rpm. MINR and MINR+5% IL had very similar 13

Table 7: Total torque loss measured on FZG test machine for the oils tested (TLC40+C14 ).

nW

200

400

1200

Load stage

MINR

MINR+IL

K1

1.34

1.30

K5

3.17

3.30

K7

6.22

4.87

K9

8.64

7.93

K1

1.55

1.53

K5

3.22

3.38

K7

6.10

4.77

K9

8.31

7.49

K1

2.22

2.19

K5

3.65

3.75

K7

6.15

4.92

K9

8.11

7.09

torque loss for all operating speeds which means that no substantial changes are found for a no-load condition. These torque losses are mainly generated by load independent losses. Figure 5b displays the torque loss for load stage K5. MINR and MINR+5% IL promoted similar experimental torque losses. However, MINR+5% IL generated slightly higher torque loss. The torque loss for the load stage K7 is displayed in Figure 5c. The MINR + 5% IL promoted a reduction in the total torque loss of around 1 Nm, which corresponds to a torque loss reduction of around 20%. The torque loss measurements for load stage K9 are presented in Figure 5d. The MINR + 5% IL promoted a reduction in torque loss between 8% and 12%, depending on the rotational speed in comparison with the mineral oil. These experimental results clearly show that the Ionic Liquid has a positive effect in the reduction of the torque loss in comparison with the original

14

Table 8: Total torque loss measured on FZG test rig for MINR with C40 gears in both test and drive gearboxes (TLC40+C40 ).

Wheel speed

Load stage

200

400

1200

15

MINR

K1

1.15

K5

3.69

K7

5.88

K9

8.88

K1

1.49

K5

3.76

K7

5.73

K9

8.53

K1

2.14

K5

4.22

K7

5.80

K9

8.08

MINR + 5% IL

TL [N.m]

TL [N.m]

MINR

10 9 8 7 6 5 4 3 2 1 0

K1

0

200

400 600 800 Rotational speed [rpm]

1000

MINR

10 9 8 7 6 5 4 3 2 1 0

1200

K5

0

200

MINR

MINR + 5% IL

K7

0

200

400 600 800 Rotational speed [rpm]

400 600 800 Rotational speed [rpm]

1000

1200

(b) load stage K5.

TL [N.m]

TL [N.m]

(a) load stage K1. 10 9 8 7 6 5 4 3 2 1 0

MINR + 5% IL

1000

MINR

10 9 8 7 6 5 4 3 2 1 0

1200

MINR + 5% IL

K9

0

(c) load stage K7.

200

400 600 800 Rotational speed [rpm]

1000

1200

(d) load stage K9.

Figure 5: Total torque loss TL in function of load and speed.

lubricant, MINR. 4.4. FZG test gearbox efficiency The calculation of the gearbox efficiency in a closed loop test rig is a function of the torque installed in the system and the torque applied by the electric motor (designated as torque loss in this work TL ). A static torque was applied on the pinion shaft TIN , as a result, the wheel shaft has a higher torque TW , function of the pinion torque and the transmission ratio (i = Z2 /Z1 ), as represented by equation 3 (see Figure 6). The wheel shaft torque values (TW ) tested were already presented in Table 6. TW = i · TIN

(3)

The torque loss (TL ), that is also the torque applied by the electric motor to keep a constant operating speed, was measured on the wheel shaft. Thus, the global efficiency of the test rig is given by equation 4. 16

TW

MOTOR TL , n 2

TIN, n1 Figure 6: Applied torque on the pinion (TIN ), wheel (TW ) and torque loss (TL ) on the FZG machine.

ηglobal =

T W − TL TW

(4)

After that, the calculation of the efficiency of the test gearbox is according the Equation (5), where the slave gearbox efficiency was determined with results presented in Table 8 and yet discussed in [8]. ηglobal ηtest = √ × 100 ηslave

(5)

Table 9 displays the efficiency values calculated for the test gearbox for both lubricants tested. Mixing MINR with 5% IL resulted in a efficiency increase of more than 0.5% at load stage K7 and more than 0.25% at load stage K9. This is a very interesting result for a single stage. The load and speed conditions of K7 and K9 are very similar to those found in first and second stage, respectively, in a real wind turbine gearbox with 2.5 MW of capacity [43]. The experimental results clearly show that some efficiency improvement, accordingly with the friction results obtained in [40], is feasible with the addition of ionic liquid to the MINR lubricant. It is quite difficult to quantify the actual value of the efficiency improvement in a real wind turbine gearbox. However, it is reasonable to assume that even if 0.25% of overall efficiency improvement 17

Table 9: Test gearbox efficiency (%) of the FZG machine for each test condition.

Wheel speed

200

400

1200

Load stage

MINR

MINR + 5% IL

K1

83.26

84.23

K5

98.73

98.61

K7

98.33

99.03

K9

98.69

98.92

K1

82.13

82.70

K5

98.71

98.56

K7

98.36

99.04

K9

98.74

99.00

K1

73.33

73.97

K5

98.52

98.42

K7

98.35

98.98

K9

98.73

99.05

18

Table 10: Operating conditions on the wear test performed.

Load stage

Wheel Torque

Wheel Speed

vt

Power

Fbn

Fr

pC40 H

pC14 H

(N m)

(rpm)

(m/s)

(W )

(N )

(N )

(M P a)

(M P a)

567.39

200

1.1

12064

16772

8506

1456

2462

K12

occurs, which can represent power savings of at least 6 kW in a 2.5 MW wind turbine.

5. Wear of FZG gears 5.1. Test procedure and operating conditions The wear test is performed under dip lubrication on the test gearbox and with oil injection on the drive gearbox, similarly as presented in section 4.2. The test gearbox oil bath temperature was kept constant at 100 ◦ C, using heaters and a cooling coil. The operating conditions used in the wear tests performed are displayed in Table 10. The test procedure can be summarized as follows: 1. Weight the gear pinion; Measure surface texture; 2. Mount gears; 3. Run load stage K12 at a rotational speed of 200 rpm during 48 hours; • Register the assembly working temperatures; • Continuous torque loss measurement; 4. Collect a lubricant sample for analysis; 5. Dismount gear; 6. Weight the gear pinion; Measure surface texture.

19

5.2. Torque Loss The torque loss measurement performed during the wear test is presented in Figure 7 for each lubricant. The torque loss of MINR + 5% IL is lower than the original formulation of MINR in the first 24 h, however, after 32 h the torque

TL [N.m]

loss with the MINR oil becomes lower than the measured with MINR + 5% IL.

16 15.5 15 14.5 14 13.5 13 12.5 12

MINR

MINR + 5% IL

K12 200 rpm 0

8

16 24 Time [h]

32

40

48

Figure 7: Total torque loss TL vs. Rotational speed for load stage K12 during wear test.

5.3. Mass loss The mass loss results during the power loss tests presented in section 4 as well as the wear tests are presented in Table 11. The accuracy of the measurement is ± 1 mg. During the power loss test campaign, no significant differences were observed between the oils. However, for the wear test the mass loss is 17% lower for MINR + 5% IL. In a previous work [40], the wear reducing performance of the mixture with regard to the lubricant without ionic liquid was related to the presence of iron phosphates on the worn surface and a correlation between the ratio iron oxides/iron hydroxy-oxides and wear volume was found. The strong influence of temperature was also pointed out with higher ratio of iron oxides/iron hydroxyoxides at 40 ◦ C resulting in higher wear reduction of the blend with regard to the gear oil without ionic liquid. While, at 100 ◦ C the ratio of iron oxides/iron 20

Table 11: Mass loss [mg] of the pinion after the power loss and wear tests.

Number of Cycles× 1000

MINR

MINR + 5% IL

Power loss

1944

10

9

Wear

864

29

24

Total

2808

39

33

hydroxy-oxides decreased to similar levels to the gear oil without ionic liquid and then wear reduction was lower. 5.4. Surface roughness The surface evolution has been accessed using 2D and 3D surface texture evaluations with a Hommelwerk T8000 device. The 2D roughness measurements were performed in the radial direction for two teeth on the pinion and two teeth on the wheel, performing three measurements for each tooth. The roughness was measured for the new pinions, for the pinions after the power loss test and after wear test. Table 12 displays the average values of the six measurements per pinion. Regarding the wear tests, the gear lubricated with MINR showed the highest variation on the average roughness. The Ra parameter of the pinion lubricated with MINR increased 13% while for the pinion lubricated with MINR + 5% IL reduced 2 %, after wear test. 5.5. Oil analysis The protection against wear provided by MINR and MINR + 5% IL is also compared through oil analysis. The analysis of the wear particles contained in the lubricant gives quite good indication about the wear of lubricated parts, since those particles in a closed box have origin in the contacting parts. The lubricant samples, collected after each group of tests (power loss and wear), were analysed by Direct Reading Ferrography in order to measure the ferrometric parameters DL (large wear particles index) and DS (small wear

21

Table 12: Roughness parameters of the tooth surfaces in radial direction: new and after power loss and wear tests.

Oil

pinion MINR wheel

pinion MINR+IL wheel

Condition

Ra

Rq

Rz

Rmax

New

0.46

0.58

3.17

4.05

Power loss

0.35

0.47

2.68

4.18

Wear

0.39

0.53

2.84

4.47

New

0.61

0.78

4.21

4.98

Power loss

0.42

0.57

3.24

4.42

Wear

0.40

0.55

3.22

4.97

New

0.53

0.67

3.60

4.37

Power loss

0.37

0.47

2.44

2.94

Wear

0.36

0.49

2.85

4.20

New

0.53

0.68

3.79

4.46

Power loss

0.34

0.44

2.34

2.76

Wear

0.34

0.45

2.63

3.49

particles index). The values of DL and DS are then used to evaluate the concentration of wear particles index - CPUC and the severity of wear particles index - ISUC, defined by equations (6) and (7), respectively. DL + DS d

(6)

DL2 − DS 2 d2

(7)

CP U C =

ISU C =

where d is the oil sample dilution which is for all cases equal to 0.1. The CPUC index grows when the sum of the small and the large particles increase, while the ISUC index grows when the number of particles with a size greater than 5μm (DL) are present in major number than the smaller than 5 μm (DS) [44]. The results for the ferrometric analysis are resumed in Table 13. The results 22

Table 13: Ferrometric indexes.

Test

Power Loss

Wear

Oil

DL

DS

CPUC

ISUC

DL

DS

CPUC

ISUC

MINR

37.4

3.3

407

1.4×105

50.7

9.1

598

2.5×105

MINR+5%IL

20.3

6.1

264

3.7×104

10.3

1.7

120

1.0×105

clearly show that the concentration of wear particles is always larger in the tests with MINR, indicating a larger amount of particle generation. The wear severity is also larger for the MINR lubricant, indicating that the wear particles are larger than the generated with MINR+ 5% IL lubricant. Such behaviour is observed both for power loss and wear tests.

6. Conclusions The use of [BMP][NTf2 ] ionic liquid as 5 wt% additive in a mineral-based fully formulated wind turbine gear oil was analysed under different testing conditions in this work. The blend tested before under a ball-on-plate reciprocating configuration had showed slight friction reduction behavior with regard to the gear oil without ionic liquid but a clear wear reduction performance. Now, using another tribological approach and developing rolling bearing (torque loss) and FZG (power loss and wear) tests, some conclusions can be drawn: the addition of the ionic liquid to the mineral-based fully formulated gear oil promoted lower torque loss in rolling bearing tests and lower power loss in FZG tests, though the small friction reduction of the blend with regard to the non-containing IL sample verified in a previous work; although the 0.25% efficiency improvements achieved are very small, they can represent savings of 6 kW in a 2.5 MW wind turbine; and strong wear reduction performance in gears can be achieved with the addition of the ionic liquid due to the physical (at low temperature) and chemical (at high temperature) interactions of the additives (ionic liquid and pre-existing additives in the gear lubricant) with surfaces. 23

Acknoledgments The authors thank to the Ministry of Science and Innovation (Spain) and to the Foundation for the Promotion in Asturias of the Applied Scientific Research and Technology (FICYT) for supporting this research within the framework of the Research Projects WINDTRIB (DPI2010-18166) and GRUPIN14-023, respectively. The authors thank to Repsol S.A. that provided the wind turbine gear oil. The authors also thank to FCT - Funda¸ca˜o para a Ciˆencia e a Tecnologia within the project EXCL/EMS-PRO/0103/2012.

Notation and Units aH Hertzian half-width [μm] CP U C Concentration of wear particles index [-] DL Large particles index [-] DS Small particles index [-] Fa axial load [N] Fbn normal load on base pitch [N] Fr rolling bearings radial load [N] ISU C Wear severity index [-] Fr rolling bearings radial load [N] L roller element contact width [mm] Mt rolling bearing torque loss [Nmm] nW wheel rotational speed [rpm] p0 Herztian contact pressure [MPa] Ra average absolute surface roughness (DIN 4768) [μm] Rmax maximum peak-to-valley height (DIN 4768) [μm] Rq root mean square average surface roughness (DIN 4768) [μm] RX equivalent radius in rolling direction of the rolling bearing contact [m] Rz mean peak-to-valley height (DIN 4768) [μm] TL FZG total torque loss [Nm] 24

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