- Email: [email protected]
Effects of thrust washer bearing surface characteristics on planetary gear train wear
Wear 432–433 (2019) 202933
Contents lists available at ScienceDirect
Wear journal homepage: www.elsevier.com/locate/wear
Effects of thrust washer bearing surface characteristics on planetary gear train wear
T
Ellen Bergsetha,∗, Mats Henrikssonb, Senad Dizdarc, Ulf Sellgrena a
KTH Royal Institute of Technology, Machine Design, SE-100 44, Stockholm, Sweden Scania CV AB, SE-151 87, Södertälje, Sweden c Höganäs AB, R&D, SE-263 83, Höganäs, Sweden b
A R T I C LE I N FO
A B S T R A C T
Keywords: Thrust washers Planetary gears Adhesive wear Surface treatment Laser cladding
Thrust washers in spur planetary gears are placed between the planet wheel and planet carrier and act as spacers and wear pads. Metal to metal sliding contact between the planet wheel – washer – carrier causes frictional power losses that, combined with starved lubrication, may cause high contact temperatures and thermo-mechanical effects that potentially trigger thermo-elastic instabilities and excessive local wear. The planetary gear system would benefit from a low-friction interface between the washer and the planet wheel. Five washers with different surface treatments were tested in a full-scale gear rig. These tests were also replicated as closely as possible in a pin-on-disc tribometer. The following types of finishing material treatments were studied: a chemical nickel coating plus polymer on a nitro-carburised surface, a combination of nitro-carburization and solid lubricant layers, electroless deposited chemical nickel coating plus polymer, nitro-carburizing, and manganese phosphating. The frictional results indicate that tribometer tests can be used to compare and classify new washer materials. Lab scale tests show that a new experimental self-lubricating tribomaterial that was applied with laser cladding has a promising potential to increase planetary gear train robustness and service life, especially if the surface is fine grinded.
1. Introduction Thrust washers are often used in planetary gear systems, when a rolling element bearing is not justifiable either from a cost-performance perspective or due to space restraints. Fig. 1 shows a spur planetary gear train, a subsystem in the heavy-duty truck gearbox shown in Fig. 2. The washers are placed between the planet wheel and the planet carrier and act as spacers and wear pads. Washers should handle sliding contacts at high relative rotational speeds. Since the lubricant may have difficulties in reaching the rolling bearing needles and washers [1], the washer operates in various lubricant regimes. Identified washer wear mechanisms are mainly adhesive wear accompanied by fatigue wear [2]. The planetary gear in Fig. 1 is equipped with spur gears, which ideally cause no axial loads on the gear shafts. However, in the reality there are no ideally stiff gear transmissions when loaded. Accordingly, it is not possible to avoid small elastic deformations that may result in significant axial thrust loads. These loads may cause wear not only on the washers but also on adjacent components and ultimately cause a failure in gear shifting. As an example, when the planetary gear system was stressed in an accelerated life durability test, wear occurred on the
∗
front washer and made it stick to the planet wheel. Thus, the washer and planet acted as one body and penetrated the carrier. To reduce the axial loads that are caused by elastic deformations, or the consequences of those axial loads, the authors distinguish three approaches: one to change the mechanical design of the washer, to use new washer materials or to perform surface treatments. To improve the mechanical design of the washer, one must consider mechanisms that cause the axial thrust load. Wang et al. [2] analysed failure mechanisms of thrust washers in a spur planetary gear system, the damage they measured imply that meshing force not only has partial force in radial direction to balance with supporting force of bearing also has a tangential force component. This tangential force component may cause the planet wheel to incline towards one side and by this force the needle roller bearing on which the planet wheel is mounted to slant. Ulezelski et al. [3] argue that axial gear thrust loads are due a skewing of the loaded needles. This effect, which was experimentally confirmed by Bair and Winer [4], can be reduced by caged needles or a close-fitting cage that constrains the skew, self-aligning needles, or changing to journal bearings. Other industry-applied design improvements that make the system less sensitive to thrust loads are to
Corresponding author. E-mail address: [email protected] (E. Bergseth).
https://doi.org/10.1016/j.wear.2019.202933 Received 12 November 2018; Received in revised form 21 May 2019; Accepted 7 June 2019 Available online 08 June 2019 0043-1648/ © 2019 Elsevier B.V. All rights reserved.
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 1. The planetary spur gear system studied, assembled view without the ring wheel (left) and exploded (right). Output shaft and carrier act as a single solid body.
Fig. 2. Section and schematic view of a Scania GRS905 heavy-duty truck gearbox. Power flow shown for gear 3, high split, and low range.
wear. Thrust washers can be made from steel in combination with different hardening treatments, but also of bronze, polymers or other low friction materials. Attention needs to be drawn to avoid materials which act as solid lubricants but are classified as pollutant or hazardous materials such as lead and copper. Coatings on steel washers, such sintered bronze-PTFE (poly-tetrafluoroethylene), manganese phosphating or electroless deposited ones can be used to provide a self-lubricating surface layer or a sacrificial solid lubricant. The purpose of self-lubrication is to stabilise coefficient of friction and reduce wear in starved lubrication or when full film lubrication cannot be established/ maintained with reliability. An important point is to prevent insufficiency of the solid lubricant in the contact. Manganese phosphated coatings acts well as a solid lubricant layer but will be worn away, mainly serves as a catalyst of a mild running-in process. One approach to ensure supply of the solid lubricants for the whole service life of a planetary gearbox could be to apply relatively thick coatings that include solid lubricant. Powder metallurgical coatings by laser cladding are weld overlays with overlay thickness up to few millimetres and a metallurgical binding to the base metal. By using laser cladding, it is possible to create various composite coatings with metallic matrices that have improved properties by adding strengthening additives [10]. When the overlay includes solid lubricant phases in form of inclusions, the solid lubricant is releases from the inclusions and spreads on the contact surface because of the contact load and temperature. This study is part of an ongoing project to explore further possible actions to reduce wear and to find parameters, patterns and thresholds that can predict failures in planetary gear trains. The overall goal of the present study is to enhance the robustness of a planetary gear system by
replace the single washer with a pair of washers that are adjacent to each other, also called double washer configuration, might sometimes help by providing extra surface interfaces if one interface fails. Moreover, one solution is to select a washer with attachments to the adjacent machine element where relative motion causes wear, but there is not much evidence to show that it is a robust solution. Further actions are to increase the radial size of the washer, machine radial grooves to produce a hydrodynamic lift or to ensure adequate supply of oil [2]. Kim et al. [5] showed that grooved thrust washers can reduce the thermo-mechanical effects with improved surface cooling and thus reduce the probability of thermo-elastic instabilities (TEI) occurring. Significant thermo-mechanical effects are common in sliding contacts between conformal surfaces, and TEI are frequently observed in clutches, gear shifting synchronizers, and disc brakes. TEI is a sudden localization of the contact and temperature increase in focal hotspots, which are of an order in magnitude smaller than the nominal contact area. Jackson and Green [6] highlight the strong relation between thrust washer failures and the presence of TEI, which is noticeable at high sliding speed and or high contact pressure. TEI changes the dominating wear mechanism from abrasive wear to scuffing [7–9]. When a local contact is frictionally heated, i.e. a hotspot appears, thermal expansion of the heated washer material may initially enhance the hydrodynamic lift [9], but, since this makes it carry an even larger portion of the load, it may cause seizure. Additionally, washers can bend easily due to the thermoelastic deformations since they are usually thin (thickness is 2 mm in this study). The other design approaches are focused on selecting an adequate washer material or surface treatment to reduce adhesive friction and 2
Wear 432–433 (2019) 202933
E. Bergseth, et al.
finding a washer design with improved friction and wear performances compared to the original washer. Reduced friction between the planet wheel and washer can prevent the washer accidently adhering to the planet wheel. Furthermore, the friction and wear should be stable over time and insensitive to variation in the lubrication conditions. The present study aims to investigate whether tribometer testing can be used as an alternative to gearbox rig tests to classify new washer materials and surface treatments. The effects of surface characteristics on washer friction and wear robustness were tested for different lubrication conditions. This study also intends to explore laser cladding as an alternative to more traditional treatments of washer contact surfaces.
between washer and carrier 2. Relative motion between washer and planet and washer and needles, no relative motion between washer and carrier 3. Relative motion between washer and carrier, no relative motion between either washer and planet wheel or washer and needles (i.e. washer, planet, and needles act as a single solid body) 4. Relative motion or slip between washer and planet wheel, washer and needles, and washer and carrier Wang et al. [2] state that motions 1 and 2 are of interest if there is relative motion and wear between washer and planet and/or needles.
2. Planetary gear system analysis
2.2. Model-based study
The gearbox in Fig. 2 can be described as three serially connected gearboxes: the splitter, the main gearbox, and the planetary gear (also called the range unit). The splitter consists of two gear-wheel pairs, low (L) and high (H). The main gearbox consists of three gear-wheel pairs (gear 1 to 6). The crawler (C) and reverse (R) gear belong to the main gearbox. The planetary gear has the ring gear locked to the gearbox housing for low range gear or to the sun gear for high range gear. The combination of high split (H) third gear and low range gear is seen as a worst-case scenario for the washers, since high power at high speed is transmitted. The speed of the planet wheel relative to the carrier is 1430 rpm. Photos of the carrier and the planet wheel disassembled from its shaft are presented in Fig. 3. Wear on the carrier can be seen in Fig. 3a. Since the planet shaft is hollow with two holes for the lubricant to pass (hole visible in Fig. 3b), the lubricant will flow to the washers due to the rotation of the lubricated double row needle bearing. Fig. 3b also shows that the needles are cageless. Fig. 3c illustrates the needle rotation when the planet shaft is fixed. When a planetary gear set is designed with unrestrained needle bearings, Ulezelski et al. [3] state that thrust washers should be dimensioned to safety carry a load equal to the net normal load on the bearings times the coefficient of friction. For the high split (H) third gear, a safety carry load of 647 N is calculated for a coefficient of friction of 0.03 between needle and planet.
The commercial FEM package ANSYS was used to analyse the sensitivity of the system by comparing the effect from steady-state frictional heating in the sliding interfaces on the mechanical contact state with a thermo-mechanical FEM-model. A suitable amount of isoparametric parabolic elements were used simulating the maximum washer stresses with less than 5% discretization error. One fifth of the system was modelled, with axisymmetric mechanical and thermal boundary conditions at the symmetry cut surfaces, and a fixed support was defined for the carrier-to-output shaft mating surface. The purpose of the model was to compare the four different idealized relative sliding motions given above by estimating the maximum contact stresses for the four cases. The transferred heat P 90 W was calculated as follows:
P = Fa⋅vrel⋅μ w
(Eq. 1)
where Fa is the applied force in N, vrel the relative speed 2 m/s between the components, and μw the coefficient of friction 0.07 between planet wheel and washer. For motion 2, which showed the largest first principal stresses compared to the other three cases, the total frictional power was 45 W to the washer surface and 45 W to the mating planet carrier and bearing needle surfaces. Based on the relative contact area with the washer, 61% of the carrier and needle frictional power was applied on the planet wheel contact surface and 39% on the sliding surfaces of the bearing needles. In the simulation, the frictional power was distributed evenly on the sliding surfaces. For all free surfaces convective boundary conditions were defined with a film coefficient of 25 W/m2 °C N and an ambient temperature of 90 °C. The simulated motion 2 stress fields in the front washer was a maximum equivalent von Mises stress of 12.3 MPa and a maximum first principal stress of 7.63 MPa. At cyclic loading, the first principal stress may induce surface crack initiation and propagation. Parts of the simulation results are shown in Fig. 4 [11]. Fig. 4b shows the uneven surface stress distribution on the front washer when an axial force is applied on the rear planet washer.
2.1. Relative washer motion The washer motions vary; in the assembly state, the washer can rotate freely, the planet wheel will rotate irrespectively at high and low range and the washer may have a relative sliding motion with all interacting surfaces, i.e. planet wheel, carrier, and needles. The relative motions between these interacting components are not deterministic, but four principal types of washer motion can be identified: 1. Relative motion between washer and planet, no relative motion
Fig. 3. Photos of a) planet carrier, input/front side, b) planet which is supported on unrestrained needles, shaft and washer and, c) illustration how relative motion between the pinion the planet shaft will cause the needles to rotate. The inner and outer diameter of the washer is 34.3 mm and 58 mm respectively and the thickness is 2 mm. 3
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 4. FEM visualisations of the a) force of 647 N applied on the rear planet washer which is equal to an axial pressure of b) 0.38 MPa on the rear washer and c) temperature distribution of the axisymmetric part.
3. Experimental procedure The experimental procedure was to perform full-scale gearbox testing at the Scania R&D department in Södertälje, Sweden, followed by pin-on-disc tribometer tests at KTH Royal Institute of Technology in Stockholm, Sweden.
3.1. Materials and surface treatments The washers were selected based on company experience. Table 1 lists washer and disc surface treatments. Washers A were surface treated with a coating process that combines the properties of electroless nickel coating and coating with fluorocarbon. Washers B were surface treated with a process that combines nitriding/nitro-carburization and solid lubricant layers. Washers C were treated with a coating process which includes MoS2 solid lubricant and PTFE-plastic. Washers D were original steel washers treated with nitro-carburizing. Washers E were manganese phosphate coated. The disc specimens for the pin-on-disc tests were made to imitate the properties of the washers as closely as possible (Fig. 5), e.g. production process and the surface treatment were the same. Same grinding method was used for all disc except the fine ground disc F2 that consisted of a subsequent finishing step using a vertical axis knife-grinding machine (Göckel G50elT) with a 120-grit CBN-grinding wheel. The only exception was F discs that were overlay welded by laser cladding. Metal powder consumable used was an experimental Ni-based alloy with 8% of MnS, a well-known solid lubricant. These discs do not represent an existing washer but are rather a demonstration of a promising metal powder alloy and laser cladding for manufacture of weld overlays with self-lubricating properties. The studied tribological washer surfaces will further on be referred to as discs A to F2. The hardness is about 40 HRC for A, B, and D washers and discs. Washers and discs C and E have about 20% lower hardness, i.e. about
Fig. 5. Tested washers and test discs adjacent to each other, a planet wheel with needles and one washer, pin, and a laser-cladded disc. Effects of loading are apparent on the used washers.
32 HRC. For disc F1 and F2, the laser weld overlay has MnS –solid lubricant - inclusions, so the HRC value does not comply with as common HRC testing. The corresponding laser weld composition without MnS has a hardness of about 40 HRC. Table 2 presents disc 3D surface roughness measurements according to ISO 25178. These values are also representative of real washers. The selected roughness parameters are arithmetical mean height of the surface Sa, RMS height of the surface Sq and, maximum height of the surface Sz (ISO), and skewness of height distribution Ssk. As seen, discs D and E, original respectively current ones, show similar roughness with mean height parameters Sa and Sq around 1 μm and maximum parameter Sz around 7 μm. The roughness is somewhat higher for discs A, B, and C, all including electroless low friction coatings. In particular
Table 1 Washer and disc surface treatments. Code
Surface treatment ®
A
Nedox SF2
B
ANS TriboNite™
C
Nedox® FM5
D E
Nitro-carburization Mn-phosphate coating
F1
Laser cladding weld overlay (disc only)
F2
Laser cladding weld overlay (disc only)
Description Electroless deposited coating with improved hardness and resistance to wear and corrosion, low coefficient of friction and non-stick properties. This treatment was preceded by nitro-carburizing. However, this was not imitated on the test discs. Instead a nitrocarburised disc with a hard anodizing plus PTFE coating was used due to experiences gained from previously tested washers. A combination of a heat treatment (nitriding/nitro-carburization) and a low friction anti-wear coating, which is suitable for use on various steels. Electroless deposited coating, which in comparison to Nedox® SF2 allows improves dry lubrication properties, hardness and wear resistance, by a combination of molybdenum disulphide, MoS2, and polytetrafluorethylene, PTFE. Original nitro-carburised washer at present. Current reference - A surface coating process that improves scuffing load resistance, running-in, and absorption of the lubricant (STD 4291-P3-Mnph-oiled), [12,13]. Overlay welding with laser cladding with Ni-based powder alloy containing 8% MnS (solid lubricant). The overlay is hard, includes solid lubricant phase and metallurgically bonded to the base metal. This is a new experimental material. The as-laser cladded overlay surface was further grinded. As F1, but with one additional subsequent fine grinding procedure to reduce the surface roughness.
4
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Table 2 Disc 3D surface roughness parameters (ISO 25178).
Arithmetical mean height of the surface Sa [μm] RMS height of the surface Sq [μm] Maximum height of the surface Sz (ISO) [μm] Skewness of height distribution Ssk [−]
New – A
New – B
New – C
Original – D
Current – E
Future – F1
Future – F2
1.0 1.3 14.8 −0.8
1.1 1.4 9.8 −0.09
1.6 2.0 17.1 0.40
0.9 1.1 7.4 −0.20
0.7 0.9 6.7 0.06
0.8 1.0 7.7 −0.06
0.6 0.7 4.5 −0.08
3.3. Tribometer testing
the maximum height parameter Sz, which is higher compared to D and E discs, between 30% and almost 300%. Disc F1, laser cladded and ground, has roughness similar to D and E discs. Disc F2, fine ground, resulted in the lowest surface roughness. Skewness of height distribution Ssk, is close to 0 (i.e. the surface profile distribution balanced ridges and valleys) for discs B, D, E, F1, and F2. In contrast, disc C have moderately over-represented ridges, while disc A has strongly overrepresented valleys (or surface pores). Surface roughness results presented here will change during the running-in. Therefore, their influence on the lubrication mode cannot be assessed based on these parameters.
The pin-on-disc tribometer shown in Fig. 9 was used to study sliding friction and wear for five washer materials/coatings, and one self-lubricating weld overlay. The tribometer has a load arm pressing the test pin onto the test disc driven by an electrical motor. The frictional force is measured by using a load cell transducer. The disc represents the washer. The pin, with a tip radius of 5 mm, was made from standard EN 100Cr6 rolling bearing steel (750 HV ∼ 60 HRC) which can represent a planet wheel material. The test pins were cleaned in an ultrasonic bath with heptane for 10 min; washed with methanol, dried in an oven and then stored in a desiccator with silica gel until testing. Test discs were used as manufactured. Tests were conducted on each test combination at a constant sliding speed of 0.5 m/s and 5000 load cycles of the disc. The humidity was 40 ± 2% and the normal load 7.15 N, which gives a nominal maximum Hertzian contact pressure of 870 MPa. The test lubricant, the same as the one used in the gear rig tests, was the commercial synthetic GL-5 oil, with a kinematic viscosity of 64 CSt at 40 °C and 11.8 CSt at 100 °C, and a density of 837 kg/m3 at 15 °C. A lubricant act as a coolant. Lubrication will thus reduce the thermal effects that Yu and Sadeghi [9] demonstrated to be of great importance. To determine the impact of lubricant conditions, the lab experiment was performed at three test set-ups (I to III). In set-up I, the oil lubricant was applied with a brush and pressurised air. A small volume of lubricant was added every 30 s. In set-up II, initial oil lubrication implies that the disc is fully covered by the lubricant before the start of the test. In set-up III, there was no oil lubrication, i.e. a dry condition. Three tests were repeated for each studied disc surface. If severe wear and/or a significant friction increase was observed at setup I or II, the disc surface was either not tested at set-up III or just tested once. Immediately after each test, worn pins and discs were examined with an optical microscope. Disc wear tracks were also measured with a stylus profilometer, a Taylor Hobson Form Talysurf PGI 800 with a stylus tip radius of 2 μm, and a SEM. The pin specific wear coefficient k in mm3/N·m was calculated as follows:
3.2. Full-scale gear rig testing The test rig shown in Fig. 6 is designed for fatigue testing, lifetime testing, and acoustic testing. Electrical motors are used to create necessary torque and in order to perform fatigue testing; a high torque capacity is needed. The maximum input torque is 13 kNm and maximum output torque is 52 kNm. In the full-scale gearbox testing all washers studied passed the seizure tests, which were performed at 80 °C with a stepwise load increase, inspection in-between each 15-min load profile. The test cycle procedure, for accelerated full-scale life durability tests at 80 °C, takes 2 h and consist of 21 steps starting at low split fourth gear (4 L) with output torque 8.0 kNm and ending at low split first gear (1 L) at output torque 20.6 kNm (Fig. 2). Input speed is constant at 1300 rpm. The criterion is to survive 48 h without any sign of seizure on planets or washers. Fig. 7 shows no or mild wear on washers facing the carrier after life durability testing, but a distinct wear band (area 1) on the front washer facing the planet means the washer did not pass the life durability test, see a summary of the results in the Appendix. All front washers mating with the planet wheel (area 1. in Fig. 7) were studied using a scanning electron microscope (SEM), see Fig. 8. Washer C is the only washer which had coating remaining in area 1. Washers C and E passed the life durability criteria. However, SEM images show adhesive wear accompanied by fatigue wear on all surfaces, although to a different extent.
k=
V sF
(Eq. 2) 3
where V is the volume in mm removed, s the sliding distance in m, and F the normal load in N. Since the number of cycles for the discs was kept constant, the sliding distance varied, because tracks were run at a different radius on the disc. 4. Results 4.1. Friction and wear behaviour Surface topography measurements of disc wear tracks from each combination (except disc A) used in set-up II are shown in Fig. 10, with the same scaling for all specimens except for discs F1, which includes few surface pores, which could be air bubbles left from the laser cladding process but also opened MnS inclusions. It is hardly possible to differ between the former and later unless one by one surface pore is investigated under a high magnification. Note that the F1 discs were ground as little as possible in order to allow for several tests runs after subsequent regrinding. Because of it, the surface of disc F2 clearly
Fig. 6. A photo of the full-scale gear-rig with microphones for acoustic testing. 5
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 7. Photos showing both sides of the front (F) and rear (B) coated washer B (ANS TriboNite™). The numbered areas indicate: 1. – The washer surface area interfacing with the planet and 2. – The washer surface area interfacing with the needles.
more uniformly. At set-up III, i.e. dry conditions, all combinations in Fig. 11 show changed wear mechanisms compared to set-up II:
shows the empty space between the two laser-cladded strings. However, the wear tracks do not include these features. The grinding ridges can still be seen for all surfaces except on disc C. Discs B, C, and E have roughly the same wear depth, about 5–10 μm. Wear tracks measured by the profilometer negligible on discs D, F1 and F2 in all cases (Fig. 10). Pin and disc wear images for set-ups II and III (B, C, E, and F treatments) are shown in Fig. 11. All pins at lubricated condition in setup II show a mild wear, with the pin wear scar edges well defined. The F1 pin has MnS inclusions on its front edge. The counter bodies (i.e. the discs) at set-up II indicate wear but to different extent. For B, the coating was almost completely worn off. Disc C has only a few shiny metal spots exposed, the MnP coating on disc E has been worn off from the highest asperities. When comparing the new experimental material, but with different grinding methods, Fig. 11 shows that disc F1 has a few wear scratches, whilst F2 has smoothened the tips of the asperities
• Pin B has mild adhesive wear and mild ploughing scratches. • Pin C disc shows severe adhesive wear and material transfer. • Pin E has mild abrasive wear with particles leftovers. • Pin F has MnS particles inclusions in the front edge. • When the pin is tested against the finer ground disc F2, the wear
decreases. Note the surface film on the pin run against disc F2 which is not seen when pin is run against disc F1 or when F2 is run in setup II. Disc F2 only shows shallow scratches, and the MnS inclusions are clearly visible as dark grey dots; but cracks adjacent to these dots existed before testing.
Fig. 8. SEM washer images after life durability gear rig tests. 6
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 9. Pin-on-disc a) illustration and b) a photo of set-up I continuous lubrication using a syringe with a brush to apply the lubricant.
As shown in Fig. 14, the coefficient of friction fluctuated heavily for all surface combinations, except for surfaces F1 and F2, which showed significantly less fluctuation. Note that the decreased roughness of surface F2 compared to F1 did not change the coefficient of friction in lubricated conditions. In dry conditions, the coefficient of friction rises gradually for all surface combinations except for F2, for which the coefficient of friction reaches a steady state value after roughly 100 s (500 cycles), see Fig. 15. The wear track width was large for the lasercladded (rough grinding) disc run at set-up III. Fig. 11 also shows that the contact has crossed the welding edges. Fine grinding of the lasercladded surface showed a big improvement in the level for the coefficient of friction, the stability over time, and a significantly reduced fluctuation. Note that C also behaves almost as stable as F2, although the fluctuation is significantly larger for C than for F2.
The effects of the changed wear mechanisms are not captured by the specific pin wear coefficient (Appendix). EDXS analysis of the F1 and F2 disc surfaces confirms the presence of Mn and S elements, Fig. 12 shows F2 disc. Due to plastic scratching wear, the coating has been released from the disc B surface, which is visible in the SEM images in Fig. 13. The EDXS analysis also showed that disc B had an increased amount of Fe on the asperities in the wear track. For disc C, the material is displaced or smeared rather than removed, which is also shown in Figs. 11 and 13. Disc C had an increased amount of fluoride in the wear track, and disc E had only fractions of coating left in the track. The phosphate crystalline morphology can be seen outside the track in Fig. 13. The SEM images also show that the coating is adhesively released from the surface using a tilting angle to better visualise the topography (the white areas are non-ferrous material or errors). Washer E showed no Mn left in the track, or on the edge, but P on the edge. No SEM image was taken on the wear tracks on discs F1 and F2.
Fig. 10. 3D surface images from measured discs used in set-up II, 4 × 4 mm area scaled in depths (note that disc F1 have a different scale). 7
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 11. Pin and disc wear photos for set-up II and III (dry). Disc materials B, C, E, F1, and F2.
5. Discussion
considered. If this happens, there will be no relative motion between washer and needles, and the washer will block the flow of the lubricant (Fig. 3) to the other side of the washer facing the carrier, which will cause that mating interface to seize up. Generally speaking, the washer wear is plasticity dominated wear. The wear seems to start with mild adhesive wear and continues with asperity sharing and plastic deformation which result in delamination wear. However the cracks evident on washers A, B and C could also origin in thermal expansion and imperfect coating deposition technique manifesting in poor coating adhesion strength to the base metal and residual stresses/process cracks. The surfaces treatments do not include any hard phases that will cause abrasive wear. Micrometer-fine and hard detached wear particles can produce some scratches but these will be soon removed by the adhesive wear mechanisms. The washer wear may manifest as noise and vibration that can be registered by the truck driver and the chances of reaching a service station before gear shifting function fails is large.
5.1. Pin-on-disc tribometer testing versus full-scale gearbox rig testing Three lubrication conditions were tested to ensure reaching similar operating conditions as in the full-scale tests. It should be noted that the pin on disc does not fully represent the washer-planet contact conditions. However, since the motions taking place between the washer and planet varies and operates in various lubricant regimes, parameters can be controlled by using the pin on disc. The pin on disc might catch phenomena, which takes place very locally on the washer. For example, the pin-on-disc tribometer tests replicated the adhesive wear mechanism in dry conditions seen on washers tested in the full-scale gearbox test rig, which indicates that the real system reaches a threshold when stressed in accelerated life durability tests. Thus, the likelihood of the washer adhering to the planet gear must be 8
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 12. SEM image (1500 X) and chemical analysis of MnS on the F2 disc.
crack propagation. According to Wu et al. [15], the wear location [15] is consistent with the location of the von-Mises stress peak. This can also be observed in the results from the model-based simulations of this gear train; the peak was always on the edge of the front washer. Modelbased simulations bridge and explain the differences between the fullscale gearbox rig tests and the tribometer tests. Washer life is hard to predict in any controlled lab tests, since 5000 cycles present less than 1% of a washer life, and surface fatigue occurs at the same time as effects from wear become significant. But, similar recommendations can made from analysis of gearbox rig tests and from lubricated tribometer tests. As washer surfaces C and E passed the gearbox rig durability tests, they also showed the most promising wear characteristics in this study. However, C is more robust, since it is
Häggström et al. [14], who focused on the reliability of the gear shifting synchronizers, suggests that the surface temperature is focused to local hot spots as the average surface temperature increase, due to coupled thermo-mechanical phenomena. This may also be true for conformal contacts in planetary gear train systems, especially for situations with large frictional power. Since the pin-on-disc does not use cyclic loading, fatigue wear was not observed in the tribometer tests. Durability tests with cyclic loading in the gearbox rig caused net-shaped crack patterns on the washer surfaces, as can be seen in Fig. 8. Cracks indicate high local contact stresses or cyclic thermo-mechanically induced stresses. The axisymmetric FEM model showed unsymmetrical stress distribution on the front washer (Fig. 4) as well as high tensile stresses, which may cause
Fig. 13. Representative SEM images (500 X with 65° tilt angle) of test discs for set-up III (dry). Arrows indicates sliding directions. 9
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 14. Representative frictional behaviours at set-up II. Note the different time scales.
In dry conditions, results show mild abrasive wear of disc E and mainly adhesive wear of discs C and B. Among the traditional washer coatings, disc C is less sensitive to the higher frictional heat power in set-up III, and the coefficient of friction is also much more stable compared to the other commercially coated discs (Fig. 15). Disc C shows smeared material rather than removed material (Fig. 11). Disc E shows a significant increase in the coefficient of friction and disc wear when running in dry conditions. This indicates that the mild wear process has been replaced with plastic scratching wear with small surface flaking as a result (Fig. 13). The chemical process to add manganese phosphate coating may, thus, not bring any major benefits in the dry case. Surface roughness amplitude due to manufacturing process seems to have a minor impact on wear and average friction properties for the tested commercially surface treated washers (Table 2). Discs C and E were both wear resistant in set-up II, but they had a about 20% lower hardness than the other discs. Frictional layers, thus, seem to be more important than having a slightly harder surface.
significantly less sensitive to an absence of lubrication.
5.2. Effects of surface characteristics on washer frictional and wear robustness for different lubrication conditions At lubricated conditions, i.e. set-ups I or II, wear is mild on all surfaces. All surface tracks on the discs tested are smoothed, and the tips of the asperities have been worn off. Hence, the frictional heat is focused to local clusters of asperities in contact. However, in dry conditions, a change in wear mechanism takes place. Friction behaviour and changes in friction level provide information on a change in the wear mechanism. Compared to the results for set-up I, the coefficient of friction in setup II does not increase for the A, B, and C discs, but it does for the D disc, which has no surface coating; see the summary of quantitative results in Appendix. That is, all types of coatings studied have reduced the coefficient of friction, and consequently also reduced the frictional heat power. The original washer material, i.e. disc D, had wear characteristics like the others in set-up I, but the frictional level and the disc wear increased significantly in set-up II. Washer D showed large adhesive wear in the full-scale rig tests (Fig. 8), but not in the tribometer tests. When the B coating was worn off (Fig. 11) the wear looked similar to the uncoated D disc (Fig. 13), but the level of the coefficient of friction was lower. For A surface material, the substrate was most likely not appropriate, since the hardened nitro-carburised surface layer is not as reactive as a non-hardened surface (the authors found no purpose in running discs A dry). Wear on disc C did not show any significant difference between set-ups I and II.
5.3. Laser cladding versus traditional treatments of washer contact surfaces In lubricated conditions, both discs F1 and F2 have a frictional and wear behaviour that is close to disc C. However, in dry conditions, discs F1 and F2 are clearly different. Disc F2 with fine grinding has succeeded in building a tribofilm, which was not generated on the rougher ground disc F1 in dry conditions. This could be due to the discs being gently dried after the fine grinding and used directly in the test. By this the solid lubricant film, already formed during grinding, could be up to a 10
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 15. Representative frictional behaviours at set-up III, i.e. dry condition. Note the different time scales.
7. Future work
certain level pre-established before the tribometer test. Moreover, the wear tracks crossing the weld joint between the strings seems to have no influence on the self-lubricating effect.
The novel surface material F1 (ground) and F2 (fine ground) may be laser-cladded directly onto the carrier surface of the main planetary gear. Such a solution could potentially make it possible to completely remove, for example, the front washer, and thus reduce the part count in the planetary gear train. Such a modification requires further industrialization of the cladding process in order to adapt it for efficient mass production. Further research on the detailed effects of surface topography, material composition of the laser cladded material and the composition of the additive package in the lubricants on the tribological properties and performance are preferably to be studied with tribometer testing. Daniel Häggström et al. [14] showed that there is a local contact surface temperature threshold which determines whether the gear shifting synchronizer in a heavy-duty truck gearbox fails after just a few gear shifting cycles or has a very long service life. This threshold is determined by when the thermo-mechanical effects will significantly affect the properties of the sliding contact, i.e. when the effects of thermo-elastic instabilities are observable. The threshold local surface temperature, and how it can be predicted, was researched, and a modelbased method that can be integrated into the gearbox management system to determine the surface temperature in real time was also proposed in Ref. [14]. Based on test results from the full-scale gearbox rig, the core model in the method presented in Ref. [14] may be adapted to monitor the risk of washer failure in the planetary gear train. Implementation of this method could facilitate accelerated testing in the full-scale gearbox rig, be implemented in the gear shifting management system in service operation, and also be implemented as a component in the condition monitoring system of an autonomous truck.
6. Conclusions The wear and frictional behaviour of five different commercial (or traditional) washer materials/coatings and one new weld overlay material were simulated for different lubrication conditions in a pin-ondisc tribometer. The results were compared with the results of full-scale gearbox rig tests with focus on wear mechanisms. The following conclusions can be drawn from this study:
• Washers that passed the full-scale tests also showed the most pro-
•
•
mising wear characteristics in lab-scaled tribotests. With sufficient lubrication, the dominating wear mechanism is mild adhesive for all combinations. Thermo-mechanical effects, and potentially also thermo-elastic instabilities, may be significant in accelerated fullscale gearbox rig tests. Such conditions cannot be studied in a tribometer. In dry conditions wear differed for the surface treatments studied: either the coating was to a large extent removed, which was reflected by a monotonic increase in the coefficient of friction, or the coating was transformed and there was a change in wear mechanism which resulted in a frictional behaviour that was more stable over time. In dry conditions the roughness of the laser-cladded material had a significant positive impact on wear and friction; the rougher surface did not form a lubrication film. The new self-lubricating material studied that was applied with laser cladding, i.e. an additive manufacturing method, seems to be a robust washer contact surface material with respect to the lubrication condition. The new surface material has a great potential to increase the life durability of planetary gear trains, especially when the cladded surface is fine ground.
Acknowledgements This work was funded by the Swedish Foundation for Strategic Research (SSF), grant number SM15-0025. We would also like to thank 11
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Scania CV AB for the opportunity to carry out this work, Bodycote for providing surface treatments, and Höganäs AB for providing specimens.
Appendix Table A1 summarizes the pin-on-disc (mean values with standard deviations in brackets) and gearbox rig test results. All lab tests presented in the table were repeated three times, except for D in set-up II and B and F1 in set-up III indicated as italic numbers. Table A1 Results from all gear rig and pin-on-disc lab tests (set-up I, II, and III). Italic numbers indicate that less than three repetitions were made, thus standard deviation is not relevant.
Gear rig
Lab test
Continuous lubrication Set-up I
Initial lubrication Set-up II
Dry Set-up III
∗ ∗∗
New – A
New – B
New – C
Original – D
Current – E
Future – F1
Future – F2
Surface treatment (steel type)
Nitro-carburised Nedox® SF2 (1265-11∗)
Nitro-carburised ANS TriboNite™ (1265-11∗)
Nitro-carburised (1265-11∗)
MnPh (Spring steel)
–
–
Seizure test P/F Life durability P/F Surface treatment (steel type)
P
P
Nedox® FM5 (126511∗) P
P
P
–
–
F
F
P
F
P
–
– Laser cladding Höganäs experimental Nibased grade with 8% MnS (weldable steel) –
Mean friction [−] Mean disc wear track [mm] Mean pin wear k [mm3/Nm] Mean friction [−] Mean disc wear track [mm] Mean pin wear k [mm3/Nm] Mean friction [−] Mean disc wear track [mm] Mean pin wear k [mm3/Nm]
®
Nitro-carburised Teflon (2541∗∗)
Nitro-carburised ANS TriboNite™ (2541∗∗)
Nedox FM5 (2541∗∗)
Nitro-carburised (2541∗∗)
MnPh (2541∗∗)
0.056 (0.002)
0.093 (0.004)
0.11 (0.009)
–
0.64 (0.048)
0.58 (0.100)
0.085 (0.002) 0.54 (0.103)
Laser cladding Höganäs experimental Nibased grade with 8% MnS (weldable steel) –
0.53 (0.253)
–
–
–
7.0 · 10−9 (2.6 10−9)
3.4 · 10−8 (4.4 10−9)
1.1 · 10−8 (6.3 10−9)
–
–
–
0.051 (0.029)
0.092 (0.002)
0.140
0.081 (0.020)
0.082 (0.010)
0.60 (0.081)
0.88 (0.119)
0.095 (0.008) 0.44 (0.065)
0.46 (0.033)
0.13 (0.01)
7.4 · 10−7 (3.1 10−8)
3.9 · 10−7 (6.8 10−8)
–
0.17
–
1.04
–
3.6 · 10−7
1.5 · 10−8 (5.0 10−9) 0.087 (0.002) 0.53 (0.253)
1.1
7.1 · 10−8 1.3· 10−6 (4.7 10−8) 0.16 – (0.032) 0.49 – (0.20)
4.6 · 10−8 1.2 · 10−7 (1.4 10−7) (1.3 10−9) 0.24 0.33 (0.14) 0.88 1.4 (0.20)
1.7 · 10−7 – (1.5 10−8)
1.0 · 10−6 2.6 · 10−6 (7.6 10−7)
1.2 · 10−8 (5.3 10−10)
0.14 (0.083) 0.55 (0.031)
1.7 · 10−8 (1.2 10−6)
Steel EN C10 (1.0301). Steel EN 34CrNiMo6, W–No. 1.6582.
References [9] [1] S. Dhar, A. Rawool, S. Ray, F. Shi, Lubricant flow and temperature prediction in a planetary gearset, SAE Technical Paper, 2011. https://doi.org/10.4271/2011-01-1235. [2] H. Wang, B. Ma, L. Niu, C. Zheng, The wear characteristics of thrust washer in conflux planetary gear train of power shift transmission, Eng. Fail. Anal. 28 (2013) 318–327 https://doi.org/10.1016/j.engfailanal.2012.06.008. [3] J.C. Ulezelski, D.G. Evans, R.J. Haka, J.D. Malloy, Needle Bearing Axial Thrust Study, (1983) SAE technical paper series. 830568. [4] S. Blair, W. Winer, Tribological characteristics of needle bearings, Tribol. 14 (1989) 269–276 https://doi.org/10.1016/S0167-8922(08)70203-7. [5] D. Kim, R.L. Jackson, I. Green, Experimental investigation of thermal and hydrodynamic effects on radially grooved thrust washer bearings, Tribol. Trans. 49 (2006) 2 https://doi. org/10.1080/05698190500544668 192–201. [6] R.L. Jackson, I. Green, The thermoelastic behavior of thrust washer bearings considering mixed lubrication, asperity contact, and thermoviscous effects, Tribol. Trans. 51 (2008) 19–32 https://doi.org/10.1080/10402000701739271. [7] R.L. Jackson, I. Green, Study of the tribological behavior of a thrust washer bearing, Tribol. Trans. 44 (3) (2001) 504–508 https://doi.org/10.1080/10402000108982489. [8] R.L. Jackson, I. Green, Experimental analysis of the wear, life and behavior of PTFE
[10]
[11]
[12] [13]
[14]
[15]
12
coated thrust washer bearings under non-axisymmetric loading, Tribol. Trans. 46 (4) (2003) 600–607 https://doi.org/10.1080/10402000308982670. T. Yu, F. Sadeghi, Thermal effects in thrust washer lubrication, ASME. J. Tribol. 124 (1) (2001) 166–177 https://doi.org/10.1115/1.1399053. I. Smurov, M. Doubenskaia, S.N. Grigoriev, D.V. Kotoban, P.A. Podrabinnik, Application of laser surface engineering to solve tribological problems, J. Frict. Wear 35 (6) (2014) 470–476 https://doi.org/10.3103/S1068366614060129. J. Lavergren, P. Suoranta, H. Wranér, Model-based Analysis of Thrust Washer Bearings in Heavy Trucks, Batchelor Thesis KTH Royal Institute of Technology, Sweden, 2016 (Swedish). J. Perry, T.S. Eyre, The effect of phosphating on the friction and wear properties of grey cast iron, Wear 43 (2) (1977) 185–197 https://doi.org/10.1016/0043-1648(77)90113-2. A. Kozlowski, W. Czechowski, Wear resistance of manganese phosphate coatings, Electrodepos. Surf. Treat. 3 (1) (1975) 55–63 https://doi.org/10.1016/0300-9416(75) 90011-5. D. Häggström, U. Sellgren, S. Björklund, Evaluation of synchronizer loading parameters and their ability to predict failure, Proc. Inst. Mech. Eng. 232 (9) (2018) 1093–1104 https://doi.org/10.1177/1350650117738395. H. Wu, G. Dong, L. Qin, W. Yuan, J. Zhang, G. Dong, Modification of spider gear back to uniform the stress and improve the anti-wear performance of a real thrust washer, Eng. Fail. Anal. 60 (2016) 107–116 https://doi.org/10.1016/j.engfailanal.2015.11.047.
Contents lists available at ScienceDirect
Wear journal homepage: www.elsevier.com/locate/wear
Effects of thrust washer bearing surface characteristics on planetary gear train wear
T
Ellen Bergsetha,∗, Mats Henrikssonb, Senad Dizdarc, Ulf Sellgrena a
KTH Royal Institute of Technology, Machine Design, SE-100 44, Stockholm, Sweden Scania CV AB, SE-151 87, Södertälje, Sweden c Höganäs AB, R&D, SE-263 83, Höganäs, Sweden b
A R T I C LE I N FO
A B S T R A C T
Keywords: Thrust washers Planetary gears Adhesive wear Surface treatment Laser cladding
Thrust washers in spur planetary gears are placed between the planet wheel and planet carrier and act as spacers and wear pads. Metal to metal sliding contact between the planet wheel – washer – carrier causes frictional power losses that, combined with starved lubrication, may cause high contact temperatures and thermo-mechanical effects that potentially trigger thermo-elastic instabilities and excessive local wear. The planetary gear system would benefit from a low-friction interface between the washer and the planet wheel. Five washers with different surface treatments were tested in a full-scale gear rig. These tests were also replicated as closely as possible in a pin-on-disc tribometer. The following types of finishing material treatments were studied: a chemical nickel coating plus polymer on a nitro-carburised surface, a combination of nitro-carburization and solid lubricant layers, electroless deposited chemical nickel coating plus polymer, nitro-carburizing, and manganese phosphating. The frictional results indicate that tribometer tests can be used to compare and classify new washer materials. Lab scale tests show that a new experimental self-lubricating tribomaterial that was applied with laser cladding has a promising potential to increase planetary gear train robustness and service life, especially if the surface is fine grinded.
1. Introduction Thrust washers are often used in planetary gear systems, when a rolling element bearing is not justifiable either from a cost-performance perspective or due to space restraints. Fig. 1 shows a spur planetary gear train, a subsystem in the heavy-duty truck gearbox shown in Fig. 2. The washers are placed between the planet wheel and the planet carrier and act as spacers and wear pads. Washers should handle sliding contacts at high relative rotational speeds. Since the lubricant may have difficulties in reaching the rolling bearing needles and washers [1], the washer operates in various lubricant regimes. Identified washer wear mechanisms are mainly adhesive wear accompanied by fatigue wear [2]. The planetary gear in Fig. 1 is equipped with spur gears, which ideally cause no axial loads on the gear shafts. However, in the reality there are no ideally stiff gear transmissions when loaded. Accordingly, it is not possible to avoid small elastic deformations that may result in significant axial thrust loads. These loads may cause wear not only on the washers but also on adjacent components and ultimately cause a failure in gear shifting. As an example, when the planetary gear system was stressed in an accelerated life durability test, wear occurred on the
∗
front washer and made it stick to the planet wheel. Thus, the washer and planet acted as one body and penetrated the carrier. To reduce the axial loads that are caused by elastic deformations, or the consequences of those axial loads, the authors distinguish three approaches: one to change the mechanical design of the washer, to use new washer materials or to perform surface treatments. To improve the mechanical design of the washer, one must consider mechanisms that cause the axial thrust load. Wang et al. [2] analysed failure mechanisms of thrust washers in a spur planetary gear system, the damage they measured imply that meshing force not only has partial force in radial direction to balance with supporting force of bearing also has a tangential force component. This tangential force component may cause the planet wheel to incline towards one side and by this force the needle roller bearing on which the planet wheel is mounted to slant. Ulezelski et al. [3] argue that axial gear thrust loads are due a skewing of the loaded needles. This effect, which was experimentally confirmed by Bair and Winer [4], can be reduced by caged needles or a close-fitting cage that constrains the skew, self-aligning needles, or changing to journal bearings. Other industry-applied design improvements that make the system less sensitive to thrust loads are to
Corresponding author. E-mail address: [email protected] (E. Bergseth).
https://doi.org/10.1016/j.wear.2019.202933 Received 12 November 2018; Received in revised form 21 May 2019; Accepted 7 June 2019 Available online 08 June 2019 0043-1648/ © 2019 Elsevier B.V. All rights reserved.
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 1. The planetary spur gear system studied, assembled view without the ring wheel (left) and exploded (right). Output shaft and carrier act as a single solid body.
Fig. 2. Section and schematic view of a Scania GRS905 heavy-duty truck gearbox. Power flow shown for gear 3, high split, and low range.
wear. Thrust washers can be made from steel in combination with different hardening treatments, but also of bronze, polymers or other low friction materials. Attention needs to be drawn to avoid materials which act as solid lubricants but are classified as pollutant or hazardous materials such as lead and copper. Coatings on steel washers, such sintered bronze-PTFE (poly-tetrafluoroethylene), manganese phosphating or electroless deposited ones can be used to provide a self-lubricating surface layer or a sacrificial solid lubricant. The purpose of self-lubrication is to stabilise coefficient of friction and reduce wear in starved lubrication or when full film lubrication cannot be established/ maintained with reliability. An important point is to prevent insufficiency of the solid lubricant in the contact. Manganese phosphated coatings acts well as a solid lubricant layer but will be worn away, mainly serves as a catalyst of a mild running-in process. One approach to ensure supply of the solid lubricants for the whole service life of a planetary gearbox could be to apply relatively thick coatings that include solid lubricant. Powder metallurgical coatings by laser cladding are weld overlays with overlay thickness up to few millimetres and a metallurgical binding to the base metal. By using laser cladding, it is possible to create various composite coatings with metallic matrices that have improved properties by adding strengthening additives [10]. When the overlay includes solid lubricant phases in form of inclusions, the solid lubricant is releases from the inclusions and spreads on the contact surface because of the contact load and temperature. This study is part of an ongoing project to explore further possible actions to reduce wear and to find parameters, patterns and thresholds that can predict failures in planetary gear trains. The overall goal of the present study is to enhance the robustness of a planetary gear system by
replace the single washer with a pair of washers that are adjacent to each other, also called double washer configuration, might sometimes help by providing extra surface interfaces if one interface fails. Moreover, one solution is to select a washer with attachments to the adjacent machine element where relative motion causes wear, but there is not much evidence to show that it is a robust solution. Further actions are to increase the radial size of the washer, machine radial grooves to produce a hydrodynamic lift or to ensure adequate supply of oil [2]. Kim et al. [5] showed that grooved thrust washers can reduce the thermo-mechanical effects with improved surface cooling and thus reduce the probability of thermo-elastic instabilities (TEI) occurring. Significant thermo-mechanical effects are common in sliding contacts between conformal surfaces, and TEI are frequently observed in clutches, gear shifting synchronizers, and disc brakes. TEI is a sudden localization of the contact and temperature increase in focal hotspots, which are of an order in magnitude smaller than the nominal contact area. Jackson and Green [6] highlight the strong relation between thrust washer failures and the presence of TEI, which is noticeable at high sliding speed and or high contact pressure. TEI changes the dominating wear mechanism from abrasive wear to scuffing [7–9]. When a local contact is frictionally heated, i.e. a hotspot appears, thermal expansion of the heated washer material may initially enhance the hydrodynamic lift [9], but, since this makes it carry an even larger portion of the load, it may cause seizure. Additionally, washers can bend easily due to the thermoelastic deformations since they are usually thin (thickness is 2 mm in this study). The other design approaches are focused on selecting an adequate washer material or surface treatment to reduce adhesive friction and 2
Wear 432–433 (2019) 202933
E. Bergseth, et al.
finding a washer design with improved friction and wear performances compared to the original washer. Reduced friction between the planet wheel and washer can prevent the washer accidently adhering to the planet wheel. Furthermore, the friction and wear should be stable over time and insensitive to variation in the lubrication conditions. The present study aims to investigate whether tribometer testing can be used as an alternative to gearbox rig tests to classify new washer materials and surface treatments. The effects of surface characteristics on washer friction and wear robustness were tested for different lubrication conditions. This study also intends to explore laser cladding as an alternative to more traditional treatments of washer contact surfaces.
between washer and carrier 2. Relative motion between washer and planet and washer and needles, no relative motion between washer and carrier 3. Relative motion between washer and carrier, no relative motion between either washer and planet wheel or washer and needles (i.e. washer, planet, and needles act as a single solid body) 4. Relative motion or slip between washer and planet wheel, washer and needles, and washer and carrier Wang et al. [2] state that motions 1 and 2 are of interest if there is relative motion and wear between washer and planet and/or needles.
2. Planetary gear system analysis
2.2. Model-based study
The gearbox in Fig. 2 can be described as three serially connected gearboxes: the splitter, the main gearbox, and the planetary gear (also called the range unit). The splitter consists of two gear-wheel pairs, low (L) and high (H). The main gearbox consists of three gear-wheel pairs (gear 1 to 6). The crawler (C) and reverse (R) gear belong to the main gearbox. The planetary gear has the ring gear locked to the gearbox housing for low range gear or to the sun gear for high range gear. The combination of high split (H) third gear and low range gear is seen as a worst-case scenario for the washers, since high power at high speed is transmitted. The speed of the planet wheel relative to the carrier is 1430 rpm. Photos of the carrier and the planet wheel disassembled from its shaft are presented in Fig. 3. Wear on the carrier can be seen in Fig. 3a. Since the planet shaft is hollow with two holes for the lubricant to pass (hole visible in Fig. 3b), the lubricant will flow to the washers due to the rotation of the lubricated double row needle bearing. Fig. 3b also shows that the needles are cageless. Fig. 3c illustrates the needle rotation when the planet shaft is fixed. When a planetary gear set is designed with unrestrained needle bearings, Ulezelski et al. [3] state that thrust washers should be dimensioned to safety carry a load equal to the net normal load on the bearings times the coefficient of friction. For the high split (H) third gear, a safety carry load of 647 N is calculated for a coefficient of friction of 0.03 between needle and planet.
The commercial FEM package ANSYS was used to analyse the sensitivity of the system by comparing the effect from steady-state frictional heating in the sliding interfaces on the mechanical contact state with a thermo-mechanical FEM-model. A suitable amount of isoparametric parabolic elements were used simulating the maximum washer stresses with less than 5% discretization error. One fifth of the system was modelled, with axisymmetric mechanical and thermal boundary conditions at the symmetry cut surfaces, and a fixed support was defined for the carrier-to-output shaft mating surface. The purpose of the model was to compare the four different idealized relative sliding motions given above by estimating the maximum contact stresses for the four cases. The transferred heat P 90 W was calculated as follows:
P = Fa⋅vrel⋅μ w
(Eq. 1)
where Fa is the applied force in N, vrel the relative speed 2 m/s between the components, and μw the coefficient of friction 0.07 between planet wheel and washer. For motion 2, which showed the largest first principal stresses compared to the other three cases, the total frictional power was 45 W to the washer surface and 45 W to the mating planet carrier and bearing needle surfaces. Based on the relative contact area with the washer, 61% of the carrier and needle frictional power was applied on the planet wheel contact surface and 39% on the sliding surfaces of the bearing needles. In the simulation, the frictional power was distributed evenly on the sliding surfaces. For all free surfaces convective boundary conditions were defined with a film coefficient of 25 W/m2 °C N and an ambient temperature of 90 °C. The simulated motion 2 stress fields in the front washer was a maximum equivalent von Mises stress of 12.3 MPa and a maximum first principal stress of 7.63 MPa. At cyclic loading, the first principal stress may induce surface crack initiation and propagation. Parts of the simulation results are shown in Fig. 4 [11]. Fig. 4b shows the uneven surface stress distribution on the front washer when an axial force is applied on the rear planet washer.
2.1. Relative washer motion The washer motions vary; in the assembly state, the washer can rotate freely, the planet wheel will rotate irrespectively at high and low range and the washer may have a relative sliding motion with all interacting surfaces, i.e. planet wheel, carrier, and needles. The relative motions between these interacting components are not deterministic, but four principal types of washer motion can be identified: 1. Relative motion between washer and planet, no relative motion
Fig. 3. Photos of a) planet carrier, input/front side, b) planet which is supported on unrestrained needles, shaft and washer and, c) illustration how relative motion between the pinion the planet shaft will cause the needles to rotate. The inner and outer diameter of the washer is 34.3 mm and 58 mm respectively and the thickness is 2 mm. 3
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 4. FEM visualisations of the a) force of 647 N applied on the rear planet washer which is equal to an axial pressure of b) 0.38 MPa on the rear washer and c) temperature distribution of the axisymmetric part.
3. Experimental procedure The experimental procedure was to perform full-scale gearbox testing at the Scania R&D department in Södertälje, Sweden, followed by pin-on-disc tribometer tests at KTH Royal Institute of Technology in Stockholm, Sweden.
3.1. Materials and surface treatments The washers were selected based on company experience. Table 1 lists washer and disc surface treatments. Washers A were surface treated with a coating process that combines the properties of electroless nickel coating and coating with fluorocarbon. Washers B were surface treated with a process that combines nitriding/nitro-carburization and solid lubricant layers. Washers C were treated with a coating process which includes MoS2 solid lubricant and PTFE-plastic. Washers D were original steel washers treated with nitro-carburizing. Washers E were manganese phosphate coated. The disc specimens for the pin-on-disc tests were made to imitate the properties of the washers as closely as possible (Fig. 5), e.g. production process and the surface treatment were the same. Same grinding method was used for all disc except the fine ground disc F2 that consisted of a subsequent finishing step using a vertical axis knife-grinding machine (Göckel G50elT) with a 120-grit CBN-grinding wheel. The only exception was F discs that were overlay welded by laser cladding. Metal powder consumable used was an experimental Ni-based alloy with 8% of MnS, a well-known solid lubricant. These discs do not represent an existing washer but are rather a demonstration of a promising metal powder alloy and laser cladding for manufacture of weld overlays with self-lubricating properties. The studied tribological washer surfaces will further on be referred to as discs A to F2. The hardness is about 40 HRC for A, B, and D washers and discs. Washers and discs C and E have about 20% lower hardness, i.e. about
Fig. 5. Tested washers and test discs adjacent to each other, a planet wheel with needles and one washer, pin, and a laser-cladded disc. Effects of loading are apparent on the used washers.
32 HRC. For disc F1 and F2, the laser weld overlay has MnS –solid lubricant - inclusions, so the HRC value does not comply with as common HRC testing. The corresponding laser weld composition without MnS has a hardness of about 40 HRC. Table 2 presents disc 3D surface roughness measurements according to ISO 25178. These values are also representative of real washers. The selected roughness parameters are arithmetical mean height of the surface Sa, RMS height of the surface Sq and, maximum height of the surface Sz (ISO), and skewness of height distribution Ssk. As seen, discs D and E, original respectively current ones, show similar roughness with mean height parameters Sa and Sq around 1 μm and maximum parameter Sz around 7 μm. The roughness is somewhat higher for discs A, B, and C, all including electroless low friction coatings. In particular
Table 1 Washer and disc surface treatments. Code
Surface treatment ®
A
Nedox SF2
B
ANS TriboNite™
C
Nedox® FM5
D E
Nitro-carburization Mn-phosphate coating
F1
Laser cladding weld overlay (disc only)
F2
Laser cladding weld overlay (disc only)
Description Electroless deposited coating with improved hardness and resistance to wear and corrosion, low coefficient of friction and non-stick properties. This treatment was preceded by nitro-carburizing. However, this was not imitated on the test discs. Instead a nitrocarburised disc with a hard anodizing plus PTFE coating was used due to experiences gained from previously tested washers. A combination of a heat treatment (nitriding/nitro-carburization) and a low friction anti-wear coating, which is suitable for use on various steels. Electroless deposited coating, which in comparison to Nedox® SF2 allows improves dry lubrication properties, hardness and wear resistance, by a combination of molybdenum disulphide, MoS2, and polytetrafluorethylene, PTFE. Original nitro-carburised washer at present. Current reference - A surface coating process that improves scuffing load resistance, running-in, and absorption of the lubricant (STD 4291-P3-Mnph-oiled), [12,13]. Overlay welding with laser cladding with Ni-based powder alloy containing 8% MnS (solid lubricant). The overlay is hard, includes solid lubricant phase and metallurgically bonded to the base metal. This is a new experimental material. The as-laser cladded overlay surface was further grinded. As F1, but with one additional subsequent fine grinding procedure to reduce the surface roughness.
4
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Table 2 Disc 3D surface roughness parameters (ISO 25178).
Arithmetical mean height of the surface Sa [μm] RMS height of the surface Sq [μm] Maximum height of the surface Sz (ISO) [μm] Skewness of height distribution Ssk [−]
New – A
New – B
New – C
Original – D
Current – E
Future – F1
Future – F2
1.0 1.3 14.8 −0.8
1.1 1.4 9.8 −0.09
1.6 2.0 17.1 0.40
0.9 1.1 7.4 −0.20
0.7 0.9 6.7 0.06
0.8 1.0 7.7 −0.06
0.6 0.7 4.5 −0.08
3.3. Tribometer testing
the maximum height parameter Sz, which is higher compared to D and E discs, between 30% and almost 300%. Disc F1, laser cladded and ground, has roughness similar to D and E discs. Disc F2, fine ground, resulted in the lowest surface roughness. Skewness of height distribution Ssk, is close to 0 (i.e. the surface profile distribution balanced ridges and valleys) for discs B, D, E, F1, and F2. In contrast, disc C have moderately over-represented ridges, while disc A has strongly overrepresented valleys (or surface pores). Surface roughness results presented here will change during the running-in. Therefore, their influence on the lubrication mode cannot be assessed based on these parameters.
The pin-on-disc tribometer shown in Fig. 9 was used to study sliding friction and wear for five washer materials/coatings, and one self-lubricating weld overlay. The tribometer has a load arm pressing the test pin onto the test disc driven by an electrical motor. The frictional force is measured by using a load cell transducer. The disc represents the washer. The pin, with a tip radius of 5 mm, was made from standard EN 100Cr6 rolling bearing steel (750 HV ∼ 60 HRC) which can represent a planet wheel material. The test pins were cleaned in an ultrasonic bath with heptane for 10 min; washed with methanol, dried in an oven and then stored in a desiccator with silica gel until testing. Test discs were used as manufactured. Tests were conducted on each test combination at a constant sliding speed of 0.5 m/s and 5000 load cycles of the disc. The humidity was 40 ± 2% and the normal load 7.15 N, which gives a nominal maximum Hertzian contact pressure of 870 MPa. The test lubricant, the same as the one used in the gear rig tests, was the commercial synthetic GL-5 oil, with a kinematic viscosity of 64 CSt at 40 °C and 11.8 CSt at 100 °C, and a density of 837 kg/m3 at 15 °C. A lubricant act as a coolant. Lubrication will thus reduce the thermal effects that Yu and Sadeghi [9] demonstrated to be of great importance. To determine the impact of lubricant conditions, the lab experiment was performed at three test set-ups (I to III). In set-up I, the oil lubricant was applied with a brush and pressurised air. A small volume of lubricant was added every 30 s. In set-up II, initial oil lubrication implies that the disc is fully covered by the lubricant before the start of the test. In set-up III, there was no oil lubrication, i.e. a dry condition. Three tests were repeated for each studied disc surface. If severe wear and/or a significant friction increase was observed at setup I or II, the disc surface was either not tested at set-up III or just tested once. Immediately after each test, worn pins and discs were examined with an optical microscope. Disc wear tracks were also measured with a stylus profilometer, a Taylor Hobson Form Talysurf PGI 800 with a stylus tip radius of 2 μm, and a SEM. The pin specific wear coefficient k in mm3/N·m was calculated as follows:
3.2. Full-scale gear rig testing The test rig shown in Fig. 6 is designed for fatigue testing, lifetime testing, and acoustic testing. Electrical motors are used to create necessary torque and in order to perform fatigue testing; a high torque capacity is needed. The maximum input torque is 13 kNm and maximum output torque is 52 kNm. In the full-scale gearbox testing all washers studied passed the seizure tests, which were performed at 80 °C with a stepwise load increase, inspection in-between each 15-min load profile. The test cycle procedure, for accelerated full-scale life durability tests at 80 °C, takes 2 h and consist of 21 steps starting at low split fourth gear (4 L) with output torque 8.0 kNm and ending at low split first gear (1 L) at output torque 20.6 kNm (Fig. 2). Input speed is constant at 1300 rpm. The criterion is to survive 48 h without any sign of seizure on planets or washers. Fig. 7 shows no or mild wear on washers facing the carrier after life durability testing, but a distinct wear band (area 1) on the front washer facing the planet means the washer did not pass the life durability test, see a summary of the results in the Appendix. All front washers mating with the planet wheel (area 1. in Fig. 7) were studied using a scanning electron microscope (SEM), see Fig. 8. Washer C is the only washer which had coating remaining in area 1. Washers C and E passed the life durability criteria. However, SEM images show adhesive wear accompanied by fatigue wear on all surfaces, although to a different extent.
k=
V sF
(Eq. 2) 3
where V is the volume in mm removed, s the sliding distance in m, and F the normal load in N. Since the number of cycles for the discs was kept constant, the sliding distance varied, because tracks were run at a different radius on the disc. 4. Results 4.1. Friction and wear behaviour Surface topography measurements of disc wear tracks from each combination (except disc A) used in set-up II are shown in Fig. 10, with the same scaling for all specimens except for discs F1, which includes few surface pores, which could be air bubbles left from the laser cladding process but also opened MnS inclusions. It is hardly possible to differ between the former and later unless one by one surface pore is investigated under a high magnification. Note that the F1 discs were ground as little as possible in order to allow for several tests runs after subsequent regrinding. Because of it, the surface of disc F2 clearly
Fig. 6. A photo of the full-scale gear-rig with microphones for acoustic testing. 5
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 7. Photos showing both sides of the front (F) and rear (B) coated washer B (ANS TriboNite™). The numbered areas indicate: 1. – The washer surface area interfacing with the planet and 2. – The washer surface area interfacing with the needles.
more uniformly. At set-up III, i.e. dry conditions, all combinations in Fig. 11 show changed wear mechanisms compared to set-up II:
shows the empty space between the two laser-cladded strings. However, the wear tracks do not include these features. The grinding ridges can still be seen for all surfaces except on disc C. Discs B, C, and E have roughly the same wear depth, about 5–10 μm. Wear tracks measured by the profilometer negligible on discs D, F1 and F2 in all cases (Fig. 10). Pin and disc wear images for set-ups II and III (B, C, E, and F treatments) are shown in Fig. 11. All pins at lubricated condition in setup II show a mild wear, with the pin wear scar edges well defined. The F1 pin has MnS inclusions on its front edge. The counter bodies (i.e. the discs) at set-up II indicate wear but to different extent. For B, the coating was almost completely worn off. Disc C has only a few shiny metal spots exposed, the MnP coating on disc E has been worn off from the highest asperities. When comparing the new experimental material, but with different grinding methods, Fig. 11 shows that disc F1 has a few wear scratches, whilst F2 has smoothened the tips of the asperities
• Pin B has mild adhesive wear and mild ploughing scratches. • Pin C disc shows severe adhesive wear and material transfer. • Pin E has mild abrasive wear with particles leftovers. • Pin F has MnS particles inclusions in the front edge. • When the pin is tested against the finer ground disc F2, the wear
decreases. Note the surface film on the pin run against disc F2 which is not seen when pin is run against disc F1 or when F2 is run in setup II. Disc F2 only shows shallow scratches, and the MnS inclusions are clearly visible as dark grey dots; but cracks adjacent to these dots existed before testing.
Fig. 8. SEM washer images after life durability gear rig tests. 6
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 9. Pin-on-disc a) illustration and b) a photo of set-up I continuous lubrication using a syringe with a brush to apply the lubricant.
As shown in Fig. 14, the coefficient of friction fluctuated heavily for all surface combinations, except for surfaces F1 and F2, which showed significantly less fluctuation. Note that the decreased roughness of surface F2 compared to F1 did not change the coefficient of friction in lubricated conditions. In dry conditions, the coefficient of friction rises gradually for all surface combinations except for F2, for which the coefficient of friction reaches a steady state value after roughly 100 s (500 cycles), see Fig. 15. The wear track width was large for the lasercladded (rough grinding) disc run at set-up III. Fig. 11 also shows that the contact has crossed the welding edges. Fine grinding of the lasercladded surface showed a big improvement in the level for the coefficient of friction, the stability over time, and a significantly reduced fluctuation. Note that C also behaves almost as stable as F2, although the fluctuation is significantly larger for C than for F2.
The effects of the changed wear mechanisms are not captured by the specific pin wear coefficient (Appendix). EDXS analysis of the F1 and F2 disc surfaces confirms the presence of Mn and S elements, Fig. 12 shows F2 disc. Due to plastic scratching wear, the coating has been released from the disc B surface, which is visible in the SEM images in Fig. 13. The EDXS analysis also showed that disc B had an increased amount of Fe on the asperities in the wear track. For disc C, the material is displaced or smeared rather than removed, which is also shown in Figs. 11 and 13. Disc C had an increased amount of fluoride in the wear track, and disc E had only fractions of coating left in the track. The phosphate crystalline morphology can be seen outside the track in Fig. 13. The SEM images also show that the coating is adhesively released from the surface using a tilting angle to better visualise the topography (the white areas are non-ferrous material or errors). Washer E showed no Mn left in the track, or on the edge, but P on the edge. No SEM image was taken on the wear tracks on discs F1 and F2.
Fig. 10. 3D surface images from measured discs used in set-up II, 4 × 4 mm area scaled in depths (note that disc F1 have a different scale). 7
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 11. Pin and disc wear photos for set-up II and III (dry). Disc materials B, C, E, F1, and F2.
5. Discussion
considered. If this happens, there will be no relative motion between washer and needles, and the washer will block the flow of the lubricant (Fig. 3) to the other side of the washer facing the carrier, which will cause that mating interface to seize up. Generally speaking, the washer wear is plasticity dominated wear. The wear seems to start with mild adhesive wear and continues with asperity sharing and plastic deformation which result in delamination wear. However the cracks evident on washers A, B and C could also origin in thermal expansion and imperfect coating deposition technique manifesting in poor coating adhesion strength to the base metal and residual stresses/process cracks. The surfaces treatments do not include any hard phases that will cause abrasive wear. Micrometer-fine and hard detached wear particles can produce some scratches but these will be soon removed by the adhesive wear mechanisms. The washer wear may manifest as noise and vibration that can be registered by the truck driver and the chances of reaching a service station before gear shifting function fails is large.
5.1. Pin-on-disc tribometer testing versus full-scale gearbox rig testing Three lubrication conditions were tested to ensure reaching similar operating conditions as in the full-scale tests. It should be noted that the pin on disc does not fully represent the washer-planet contact conditions. However, since the motions taking place between the washer and planet varies and operates in various lubricant regimes, parameters can be controlled by using the pin on disc. The pin on disc might catch phenomena, which takes place very locally on the washer. For example, the pin-on-disc tribometer tests replicated the adhesive wear mechanism in dry conditions seen on washers tested in the full-scale gearbox test rig, which indicates that the real system reaches a threshold when stressed in accelerated life durability tests. Thus, the likelihood of the washer adhering to the planet gear must be 8
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 12. SEM image (1500 X) and chemical analysis of MnS on the F2 disc.
crack propagation. According to Wu et al. [15], the wear location [15] is consistent with the location of the von-Mises stress peak. This can also be observed in the results from the model-based simulations of this gear train; the peak was always on the edge of the front washer. Modelbased simulations bridge and explain the differences between the fullscale gearbox rig tests and the tribometer tests. Washer life is hard to predict in any controlled lab tests, since 5000 cycles present less than 1% of a washer life, and surface fatigue occurs at the same time as effects from wear become significant. But, similar recommendations can made from analysis of gearbox rig tests and from lubricated tribometer tests. As washer surfaces C and E passed the gearbox rig durability tests, they also showed the most promising wear characteristics in this study. However, C is more robust, since it is
Häggström et al. [14], who focused on the reliability of the gear shifting synchronizers, suggests that the surface temperature is focused to local hot spots as the average surface temperature increase, due to coupled thermo-mechanical phenomena. This may also be true for conformal contacts in planetary gear train systems, especially for situations with large frictional power. Since the pin-on-disc does not use cyclic loading, fatigue wear was not observed in the tribometer tests. Durability tests with cyclic loading in the gearbox rig caused net-shaped crack patterns on the washer surfaces, as can be seen in Fig. 8. Cracks indicate high local contact stresses or cyclic thermo-mechanically induced stresses. The axisymmetric FEM model showed unsymmetrical stress distribution on the front washer (Fig. 4) as well as high tensile stresses, which may cause
Fig. 13. Representative SEM images (500 X with 65° tilt angle) of test discs for set-up III (dry). Arrows indicates sliding directions. 9
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 14. Representative frictional behaviours at set-up II. Note the different time scales.
In dry conditions, results show mild abrasive wear of disc E and mainly adhesive wear of discs C and B. Among the traditional washer coatings, disc C is less sensitive to the higher frictional heat power in set-up III, and the coefficient of friction is also much more stable compared to the other commercially coated discs (Fig. 15). Disc C shows smeared material rather than removed material (Fig. 11). Disc E shows a significant increase in the coefficient of friction and disc wear when running in dry conditions. This indicates that the mild wear process has been replaced with plastic scratching wear with small surface flaking as a result (Fig. 13). The chemical process to add manganese phosphate coating may, thus, not bring any major benefits in the dry case. Surface roughness amplitude due to manufacturing process seems to have a minor impact on wear and average friction properties for the tested commercially surface treated washers (Table 2). Discs C and E were both wear resistant in set-up II, but they had a about 20% lower hardness than the other discs. Frictional layers, thus, seem to be more important than having a slightly harder surface.
significantly less sensitive to an absence of lubrication.
5.2. Effects of surface characteristics on washer frictional and wear robustness for different lubrication conditions At lubricated conditions, i.e. set-ups I or II, wear is mild on all surfaces. All surface tracks on the discs tested are smoothed, and the tips of the asperities have been worn off. Hence, the frictional heat is focused to local clusters of asperities in contact. However, in dry conditions, a change in wear mechanism takes place. Friction behaviour and changes in friction level provide information on a change in the wear mechanism. Compared to the results for set-up I, the coefficient of friction in setup II does not increase for the A, B, and C discs, but it does for the D disc, which has no surface coating; see the summary of quantitative results in Appendix. That is, all types of coatings studied have reduced the coefficient of friction, and consequently also reduced the frictional heat power. The original washer material, i.e. disc D, had wear characteristics like the others in set-up I, but the frictional level and the disc wear increased significantly in set-up II. Washer D showed large adhesive wear in the full-scale rig tests (Fig. 8), but not in the tribometer tests. When the B coating was worn off (Fig. 11) the wear looked similar to the uncoated D disc (Fig. 13), but the level of the coefficient of friction was lower. For A surface material, the substrate was most likely not appropriate, since the hardened nitro-carburised surface layer is not as reactive as a non-hardened surface (the authors found no purpose in running discs A dry). Wear on disc C did not show any significant difference between set-ups I and II.
5.3. Laser cladding versus traditional treatments of washer contact surfaces In lubricated conditions, both discs F1 and F2 have a frictional and wear behaviour that is close to disc C. However, in dry conditions, discs F1 and F2 are clearly different. Disc F2 with fine grinding has succeeded in building a tribofilm, which was not generated on the rougher ground disc F1 in dry conditions. This could be due to the discs being gently dried after the fine grinding and used directly in the test. By this the solid lubricant film, already formed during grinding, could be up to a 10
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Fig. 15. Representative frictional behaviours at set-up III, i.e. dry condition. Note the different time scales.
7. Future work
certain level pre-established before the tribometer test. Moreover, the wear tracks crossing the weld joint between the strings seems to have no influence on the self-lubricating effect.
The novel surface material F1 (ground) and F2 (fine ground) may be laser-cladded directly onto the carrier surface of the main planetary gear. Such a solution could potentially make it possible to completely remove, for example, the front washer, and thus reduce the part count in the planetary gear train. Such a modification requires further industrialization of the cladding process in order to adapt it for efficient mass production. Further research on the detailed effects of surface topography, material composition of the laser cladded material and the composition of the additive package in the lubricants on the tribological properties and performance are preferably to be studied with tribometer testing. Daniel Häggström et al. [14] showed that there is a local contact surface temperature threshold which determines whether the gear shifting synchronizer in a heavy-duty truck gearbox fails after just a few gear shifting cycles or has a very long service life. This threshold is determined by when the thermo-mechanical effects will significantly affect the properties of the sliding contact, i.e. when the effects of thermo-elastic instabilities are observable. The threshold local surface temperature, and how it can be predicted, was researched, and a modelbased method that can be integrated into the gearbox management system to determine the surface temperature in real time was also proposed in Ref. [14]. Based on test results from the full-scale gearbox rig, the core model in the method presented in Ref. [14] may be adapted to monitor the risk of washer failure in the planetary gear train. Implementation of this method could facilitate accelerated testing in the full-scale gearbox rig, be implemented in the gear shifting management system in service operation, and also be implemented as a component in the condition monitoring system of an autonomous truck.
6. Conclusions The wear and frictional behaviour of five different commercial (or traditional) washer materials/coatings and one new weld overlay material were simulated for different lubrication conditions in a pin-ondisc tribometer. The results were compared with the results of full-scale gearbox rig tests with focus on wear mechanisms. The following conclusions can be drawn from this study:
• Washers that passed the full-scale tests also showed the most pro-
•
•
mising wear characteristics in lab-scaled tribotests. With sufficient lubrication, the dominating wear mechanism is mild adhesive for all combinations. Thermo-mechanical effects, and potentially also thermo-elastic instabilities, may be significant in accelerated fullscale gearbox rig tests. Such conditions cannot be studied in a tribometer. In dry conditions wear differed for the surface treatments studied: either the coating was to a large extent removed, which was reflected by a monotonic increase in the coefficient of friction, or the coating was transformed and there was a change in wear mechanism which resulted in a frictional behaviour that was more stable over time. In dry conditions the roughness of the laser-cladded material had a significant positive impact on wear and friction; the rougher surface did not form a lubrication film. The new self-lubricating material studied that was applied with laser cladding, i.e. an additive manufacturing method, seems to be a robust washer contact surface material with respect to the lubrication condition. The new surface material has a great potential to increase the life durability of planetary gear trains, especially when the cladded surface is fine ground.
Acknowledgements This work was funded by the Swedish Foundation for Strategic Research (SSF), grant number SM15-0025. We would also like to thank 11
Wear 432–433 (2019) 202933
E. Bergseth, et al.
Scania CV AB for the opportunity to carry out this work, Bodycote for providing surface treatments, and Höganäs AB for providing specimens.
Appendix Table A1 summarizes the pin-on-disc (mean values with standard deviations in brackets) and gearbox rig test results. All lab tests presented in the table were repeated three times, except for D in set-up II and B and F1 in set-up III indicated as italic numbers. Table A1 Results from all gear rig and pin-on-disc lab tests (set-up I, II, and III). Italic numbers indicate that less than three repetitions were made, thus standard deviation is not relevant.
Gear rig
Lab test
Continuous lubrication Set-up I
Initial lubrication Set-up II
Dry Set-up III
∗ ∗∗
New – A
New – B
New – C
Original – D
Current – E
Future – F1
Future – F2
Surface treatment (steel type)
Nitro-carburised Nedox® SF2 (1265-11∗)
Nitro-carburised ANS TriboNite™ (1265-11∗)
Nitro-carburised (1265-11∗)
MnPh (Spring steel)
–
–
Seizure test P/F Life durability P/F Surface treatment (steel type)
P
P
Nedox® FM5 (126511∗) P
P
P
–
–
F
F
P
F
P
–
– Laser cladding Höganäs experimental Nibased grade with 8% MnS (weldable steel) –
Mean friction [−] Mean disc wear track [mm] Mean pin wear k [mm3/Nm] Mean friction [−] Mean disc wear track [mm] Mean pin wear k [mm3/Nm] Mean friction [−] Mean disc wear track [mm] Mean pin wear k [mm3/Nm]
®
Nitro-carburised Teflon (2541∗∗)
Nitro-carburised ANS TriboNite™ (2541∗∗)
Nedox FM5 (2541∗∗)
Nitro-carburised (2541∗∗)
MnPh (2541∗∗)
0.056 (0.002)
0.093 (0.004)
0.11 (0.009)
–
0.64 (0.048)
0.58 (0.100)
0.085 (0.002) 0.54 (0.103)
Laser cladding Höganäs experimental Nibased grade with 8% MnS (weldable steel) –
0.53 (0.253)
–
–
–
7.0 · 10−9 (2.6 10−9)
3.4 · 10−8 (4.4 10−9)
1.1 · 10−8 (6.3 10−9)
–
–
–
0.051 (0.029)
0.092 (0.002)
0.140
0.081 (0.020)
0.082 (0.010)
0.60 (0.081)
0.88 (0.119)
0.095 (0.008) 0.44 (0.065)
0.46 (0.033)
0.13 (0.01)
7.4 · 10−7 (3.1 10−8)
3.9 · 10−7 (6.8 10−8)
–
0.17
–
1.04
–
3.6 · 10−7
1.5 · 10−8 (5.0 10−9) 0.087 (0.002) 0.53 (0.253)
1.1
7.1 · 10−8 1.3· 10−6 (4.7 10−8) 0.16 – (0.032) 0.49 – (0.20)
4.6 · 10−8 1.2 · 10−7 (1.4 10−7) (1.3 10−9) 0.24 0.33 (0.14) 0.88 1.4 (0.20)
1.7 · 10−7 – (1.5 10−8)
1.0 · 10−6 2.6 · 10−6 (7.6 10−7)
1.2 · 10−8 (5.3 10−10)
0.14 (0.083) 0.55 (0.031)
1.7 · 10−8 (1.2 10−6)
Steel EN C10 (1.0301). Steel EN 34CrNiMo6, W–No. 1.6582.
References [9] [1] S. Dhar, A. Rawool, S. Ray, F. Shi, Lubricant flow and temperature prediction in a planetary gearset, SAE Technical Paper, 2011. https://doi.org/10.4271/2011-01-1235. [2] H. Wang, B. Ma, L. Niu, C. Zheng, The wear characteristics of thrust washer in conflux planetary gear train of power shift transmission, Eng. Fail. Anal. 28 (2013) 318–327 https://doi.org/10.1016/j.engfailanal.2012.06.008. [3] J.C. Ulezelski, D.G. Evans, R.J. Haka, J.D. Malloy, Needle Bearing Axial Thrust Study, (1983) SAE technical paper series. 830568. [4] S. Blair, W. Winer, Tribological characteristics of needle bearings, Tribol. 14 (1989) 269–276 https://doi.org/10.1016/S0167-8922(08)70203-7. [5] D. Kim, R.L. Jackson, I. Green, Experimental investigation of thermal and hydrodynamic effects on radially grooved thrust washer bearings, Tribol. Trans. 49 (2006) 2 https://doi. org/10.1080/05698190500544668 192–201. [6] R.L. Jackson, I. Green, The thermoelastic behavior of thrust washer bearings considering mixed lubrication, asperity contact, and thermoviscous effects, Tribol. Trans. 51 (2008) 19–32 https://doi.org/10.1080/10402000701739271. [7] R.L. Jackson, I. Green, Study of the tribological behavior of a thrust washer bearing, Tribol. Trans. 44 (3) (2001) 504–508 https://doi.org/10.1080/10402000108982489. [8] R.L. Jackson, I. Green, Experimental analysis of the wear, life and behavior of PTFE
[10]
[11]
[12] [13]
[14]
[15]
12
coated thrust washer bearings under non-axisymmetric loading, Tribol. Trans. 46 (4) (2003) 600–607 https://doi.org/10.1080/10402000308982670. T. Yu, F. Sadeghi, Thermal effects in thrust washer lubrication, ASME. J. Tribol. 124 (1) (2001) 166–177 https://doi.org/10.1115/1.1399053. I. Smurov, M. Doubenskaia, S.N. Grigoriev, D.V. Kotoban, P.A. Podrabinnik, Application of laser surface engineering to solve tribological problems, J. Frict. Wear 35 (6) (2014) 470–476 https://doi.org/10.3103/S1068366614060129. J. Lavergren, P. Suoranta, H. Wranér, Model-based Analysis of Thrust Washer Bearings in Heavy Trucks, Batchelor Thesis KTH Royal Institute of Technology, Sweden, 2016 (Swedish). J. Perry, T.S. Eyre, The effect of phosphating on the friction and wear properties of grey cast iron, Wear 43 (2) (1977) 185–197 https://doi.org/10.1016/0043-1648(77)90113-2. A. Kozlowski, W. Czechowski, Wear resistance of manganese phosphate coatings, Electrodepos. Surf. Treat. 3 (1) (1975) 55–63 https://doi.org/10.1016/0300-9416(75) 90011-5. D. Häggström, U. Sellgren, S. Björklund, Evaluation of synchronizer loading parameters and their ability to predict failure, Proc. Inst. Mech. Eng. 232 (9) (2018) 1093–1104 https://doi.org/10.1177/1350650117738395. H. Wu, G. Dong, L. Qin, W. Yuan, J. Zhang, G. Dong, Modification of spider gear back to uniform the stress and improve the anti-wear performance of a real thrust washer, Eng. Fail. Anal. 60 (2016) 107–116 https://doi.org/10.1016/j.engfailanal.2015.11.047.