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Analysis of automotive transmission gearbox synchronizer wear due to torsional vibration and the parameters influencing wear reduction
Engineering Failure Analysis 105 (2019) 427–443
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Engineering Failure Analysis journal homepage: www.elsevier.com/locate/engfailanal
Analysis of automotive transmission gearbox synchronizer wear due to torsional vibration and the parameters influencing wear reduction
T
Barathiraja Ka,b, , Devaradjane Gb, Jibin Paula, Rakesh Sa, Gajanan Jamadadea ⁎
a b
Mahindra & Mahindra Ltd, Mahindra Research Valley, Chennai, India Department of Automobile Engineering, Madras Institute of Technology, Anna University, Chennai, India
ARTICLE INFO
ABSTRACT
Keywords: Manual transmission Carbon synchronizers Wear Torsional vibration Lubrication Angular acceleration
Synchronizers are the heart for the manual transmission, automated manual transmission and dual clutch transmission. The synchronizers match the speed of the target gears during gear shifts. Downsizing the high-power density engine develops higher angular accelerations. Higher angular accelerations create torsional vibrations and are detrimental to the life of the synchronizers. The synchronizer rings can move freely in the available space due to torsional vibration. The synchronizers which experience higher angular acceleration collide with the surrounding parts and wear out. The wear of synchronizer carbon liner reduces the wear gap to zero. The zerowear gap hampers the synchronizer functionality and leads to gear clash. This paper presents the impact of angular acceleration on the life of carbon synchronizer ring and the parameters which are influencing to overcome the failure are studied. A bench test set-up was developed to simulate the vehicle level angular accelerations. The bench test results show direct correlation with the life of the synchronizer on the vehicle. Torsional vibration dampening using clutch size, oil viscosity, guiding of synchronizer ring, and oil volume are studied. The synchronizer carbon liner wear reduction is studied and validated with different bench and vehicle level tests.
1. Introduction Synchronizers are conical mechanical clutches used to synchronize the speed of the rotating parts. Synchronizers are used in manual transmission, automated manual transmission, and dual clutch transmissions for passenger and commercial vehicles. The most commonly used material for synchronizer is brass. To improve the friction of cone surface, wear resistance, and durability of synchronizer, carbon synchronizers are widely used. Under severe test and operating conditions, the carbon synchronizer shows higher wear resistance and durability when compared with brass synchronizers. In vehicles, the internal combustion engines are the main source for the angular acceleration and torsional fluctuations due to the engine firing order [1,2]. Stringent emission norms and the demand for fuel efficient engines force the automotive industries to downsize the engines. Small sized high-power density internal combustion engines generate high torsional vibration [3]. The torsional fluctuation generated by the engine is transmitted to the gearbox through clutch. Dual mass flywheel can filter the torsional fluctuations produced by the engine [4]. However, generally not used in commercial vehicle engines, due to high cost when
Abbreviations: TV, Torsional vibration; GSD, Gear shift durability.; CFD, Computational fluid dynamics ⁎ Corresponding author at: Mahindra & Mahindra Ltd, Mahindra Research Valley, Chennai, India. E-mail address: [email protected] (B. K). https://doi.org/10.1016/j.engfailanal.2019.06.084 Received 8 March 2019; Received in revised form 25 June 2019; Accepted 25 June 2019 Available online 26 June 2019 1350-6307/ © 2019 Elsevier Ltd. All rights reserved.
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Fig. 1. 6 Speed gearbox layout
compared with single mass flywheel. All the fixed gearwheels in the commercial vehicle coaxial manual transmission gearbox are mounted on the countershaft. This leads to excitation of all the idler gears by the countershaft, both in neutral and with any gear step engaged condition [1]. The torsional vibration creates the brass synchronizer ring lug failure and synchronizer ring wear gap reduction [5,7]. This paper reports on an experimental investigation of carbon synchronizer liner wear reduction in a conventional manual gearbox under torsional vibration condition. First, the endurance capacity or amount of wear on the designed synchronizer are studied with the μ-comp test rig, bench level GSD test rig and vehicle GSD test. Then the vehicle is instrumented to capture the driver driving pattern such as gear usage pattern and the engine speed. The gearbox and the engine are instrumented to capture the torsional vibrations on the engine flywheel and input shaft of the gearbox after the clutch. This measurement describes the input torsional fluctuations to the gearbox. Then the gearbox is instrumented and tested on the TV test rig which replicates the actual vehicle carbon synchronizer liner wear. On the TV test rig, the gearbox is tested on different clutch sizes, oil volume, oil viscosity and inner diameter guiding of synchronizer ring. The improvements are validated on the vehicle to confirm meeting of 100,000 km target distance. 2. Statement of problem The gearbox taken for the study is inline manual transmission having input shaft, counter shaft and output shaft. It has dual cone carbon synchronizers in the 1st, 2nd, 3rd and 4th gear positions and single cone carbon synchronizer in 5th and 6th gear positions. All the synchronizers are located on the output shaft as shown in Fig. 1. The synchronizer assembly consists of synchronizer hub, synchronizer sleeve, synchronizer blocker ring, dog teeth and strut detent. The outer synchronizer ring and inner synchronizer ring are made of steel and have cone surface. The carbon liner is bonded on the steel intermediate synchronizer ring. The outer, intermediate and inner synchronizer ring parts are assembled loosely with in the available space of hub to gear assembly as shown in Figs. 2 and 3. The synchronizer assembly is designed to have a nominal wear gap of 1.5 mm. The useful life target of the gearbox is 100,000 km. The synchronizer rings wear off quickly in the vehicle validation in less than 20,000 km especially in 1st, 2nd, 3rd & 4th gear positions. Due to ring wear, the wear gap as shown in Fig. 3, is getting closed or become zero. Zero wear gap hampers synchronization and the gearbox results into early gear crash on the non-engaged synchronizer rings. 3. Experimental set-up The experimental test flow is given in Fig. 4. The synchronizer design is validated for the coefficient of friction generation and the endurance life with the help of μ-comp synchronizer endurance test, Transmission level GSD test and Vehicle level GSD. These tests do not include any external noise factors. Then the external factors which affects the synchronizer function are checked using customer driving pattern analysis, TV test and Highway durability test. The above experimental test flow can be broadly divided into two types: Bench level test and Vehicle level test according to Table 1. The bench tests are the replica of the vehicle test. The bench test gives the flexibility of controlling the test parameters and reduces the total test duration. The external noise factors are not present in the bench test. However, the vehicle tests are carried out to simulate the actual vehicle driving conditions.
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Fig. 2. Synchronizer assembly
Fig. 3. Synchronizer ring wear gap
3.1. Bench tests 3.1.1. μ-comp Synchronizer endurance test rig The component level endurance test is conducted using a conventional μ-comp test rig as shown in Fig. 5. The synchronizer hub is mounted on the main shaft of the rig. The main shaft is connected to the electric motor. The sleeve slides over the hub. A hydraulic shift actuator is used to move the sleeve. The sleeve axial force and the linear travel are measured using force sensor and linear accelerometers. The outer and the inner synchronizer rings are connected to the inertia disc through the hub assembly. The intermediate synchronizer ring is connected to the gear and it is rigidly fixed to the test rig. An RPM sensor measures the main shaft speed. The lubrication oil is supplied over the synchronizer ring through a pump. The lubrication oil flow rate is measured and controlled using a flow meter. A heater is provided in the hydraulic system to heat up the oil during test. The cone torque generated on the synchronizer cone surfaces is measured using a strain gauge. The coefficient of friction and the wear are calculated by the system. The test specification is according to Table 2. 3.1.2. Transmission GSD test rig On the GSD test rig (Fig. 6), the electric motor is coupled to the output shaft. The clutch disc is mounted on the input shaft. Two 429
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Fig. 4. Block diagram of test flow sequence Table 1 Bench tests and vehicle tests Bench tests
Vehicle tests
μ-comp synchronizer endurance test Transmission GSD Test TV test rig Engine Dyno test
Vehicle GSD test Driver driving pattern measurement Vehicle angular acceleration measurement High way durability test
Fig. 5. μ-comp test rig setup
pneumatic actuators are connected to the gear shift knob for the gear select and gear shift directions. The input and the output shaft rpm are measured using speed sensors. The gear shift forces and the travel are measured using strain gauges and position sensors. A temperature sensor is used to monitor the transmission oil temperature. For dynamic gearshift, the speed of the input shaft during gear shift is as per the Table 3. The gears are shifted in the sequence according to Table 4. 430
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Table 2 Test specification for μ-comp test Parameter
Specification
Mean cone diameter Cone angle Force Delta speed Inertia Cycle time Test cycle
110 mm 8.5° 2000 N 1500 rpm 0.125 Kgm2 3.0 s 100,000
Fig. 6. Gear shift durability test set-up Table 3 GSD test bench gear shift speeds Shift type
Input shaft speed (rpm)
Up-shift Down-shift
1800 1200
Table 4 Gearbox GSD test cycle Gear shift sequence
No of cycles
1-2-1 2-3-2 3-4-3 4-5-4 5-6-5
50,000 100,000 100,000 100,000 100,000
3.1.3. TV test rig The TV (Torsional Vibration) test rig consists of gearbox, drive motor, propeller shaft & adopter pivot as shown in Fig. 7. The clutch disc is mounted on the gearbox input shaft. The electric motor is connected to the gearbox output shaft through the coupling and the propeller shaft. The adopter plate to the mount is offset by 15°-20° on the vertical axis and 10°-15° in the horizontal axis. The propeller shaft yokes have a phase difference of 90° as shown in Fig. 8. This gives a phase difference during shaft rotation. A magnetic pickup speed sensor is used to measure the input shaft speed. The acceleration is processed using sensor electronics. The gearbox is engaged in the 5th gear position (1:1 gear ratio). The motor is continuously driven at 1100 rpm. This rpm with propeller shaft phase difference creates 2000 rad/s2 angular acceleration on the input shaft. This set-up vibrates the synchronizer ring assembly. The synchronizer life in the vehicle can be directly correlated with the TV test rig. A DOE is performed using one factor at a time DOE methodology. The test sequence on the TV test rig is as per the Table 5. 431
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Fig. 7. Torsional vibration test set-up
Fig. 8. Propeller shaft phase out Table 5 Test configuration of TV test rig Test No
Hub to ring clearance (mm)
Oil grade
Oil volume (l)
1 2 3 4
3 0.5 0.5 0.5
SAE80 SAE80 SAE80W90 SAE80W90
5 5 5 7
3.1.4. Engine Dyno test The test transmission is coupled to the engine on the engine dyno test bed as shown in Fig. 9. The gearbox is pre-selected to the 5th gear position. The engine power is transferred to the gearbox through the flywheel & clutch. The engine rpm is measured with a speed sensor on the flywheel. The engine is continuously running at 1100–1200 rpm for 300 h. The test configuration for the gearbox is as per the Table 6. 3.2. Vehicle test The gearbox is mounted on a vehicle to test the gearbox in the actual driving conditions. The vehicle is loaded to the GVW to simulate the actual load condition. 3.2.1. Vehicle GSD test In the vehicle the gears are shifted in the sequence of 1-2-3-4-5-6-5-4-3-2-1. In this test, trained drivers do quick gear shift at an 432
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Fig. 9. Engine dyno test set-up Table 6 Test configuration of engine dyno Test no
Clutch size
Hub to ring clearance (mm)
Oil grade
Oil volume (l)
1 2
Standard diameter 10% increased diameter
3 0.5
SAE80 SAE80W90
5 7
engine speed as per the Table 7. The vehicle was driven for 15,000 shift cycles. This is an abuse test for the synchronizer components. 3.2.2. Vehicle angular acceleration On the vehicle, the engine and the gearbox are instrumented to measure the angular accelerations. A magnetic pickup tachometer is mounted on the gearbox focusing the engine flywheel tooth. Another magnetic pickup tachometer was positioned in the gearbox focusing the input shaft gear. This is used to measure the speed, in revolutions per minute. Whenever the tachometer sensor detects a gear tooth, it generates a pulse. The tachometers are connected to SCADAS, and LMS Test Lab is used to measure the angular accelerations of the flywheel and the input shaft gear. 3.2.3. Vehicle high way durability test The vehicle is driven on the high-speed durability test cycle. Trained drivers are driving the vehicle on the highway. The vehicle is quickly taken into 5th gear by quick gear shifting from “N” gear. Then the vehicle is driven at an engine speed of 1100–1400 rpm on 5th gear. 4. Results & Discussion The synchronizer ring wear can occur due to poor design and or due to external factors. The synchronizer capacity is used to validate the robustness in design. The external factors which causes wear are discussed under drivetrain vibrations. 4.1. Synchronizer capacity The synchronizer capacity was validated to check the life of the synchronizer ring. This validates the capability of the rings to Table 7 Vehicle GSD gear shift speeds Shift type
Input shaft speed (rpm)
Up-shift Down-shift
1900 1900
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Table 8 Material limits of carbon synchronizer Specification
Material Limits
Energy limit Surface pressure Surface velocity
1.5 J/mm2 10 N/mm2 8 to 10 m/s
perform its intended function of synchronization. 4.1.1. Synchronizer level GSD test The synchronizer ring assembly was validated in μ-comp test rig for the endurance life. To meet the endurance life of the gearbox, the material limits of the synchronizer for each gear shift operations should be more than the synchronizer capacity as per the Table 8. The synchronizer capacity material limits are calculated using the expression [6].
Energy Limit =
1 IRef 2
Surface Pressure =
2
Fs Nc A sin
Surface Velocity = dmax
= abs ( rpm
i
= abs (Grpm
(1) (2) rpm
(3)
60
(4)
a)
(5)
Hrpm )
This test was performed in a controlled environment with no vibration in the system. The test gives a stable friction and μdynamic is 0.105 at the end stage. The total wear is 0.21 mm for the entire test cycle (Fig. 10). The wear distribution is 55:45 on the inner and outer rings with no abnormalities in the entire test cycle. When the vehicle is driven on top gear with max engine rpm, the idle synchronizer packs in the low gear experience a high surface velocity. The 1st gear experiences 30% higher surface velocity than the carbon limits. In the next test, the hub speed is increased to match the vehicle Δrpm on the test rig without performing the gear shift operation. This generates 30% higher surface velocity on the synchronizer cone surfaces. The test was conducted for 5 h and no wear observed. 4.1.2. Transmission level GSD Endurance test The wear of 1st synchronizer ring in the 1-2-1 gear shift is 0.55 mm. The 1st gear reflected inertia is higher than other gears. The
Fig. 10. μ-comp coefficient of friction 434
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Wear Gap (mm)
B. K, et al.
1.8 1.6 1.4 1.2 1 0.8 0.6 0.4 0.2 0 1
2
3
4
5
6
Gear Wear Gap Before Test
Wear Gap A!er Test
Fig. 11. Wear gap in transmission GSD
higher inertia causes higher wear on the synchronizer rings when compared to other gear positions. During upshift the gearbox drag helps for synchronization. In downshift the gearbox has to work against the drag for synchronization. The higher work done by the synchronizer by accelerating the speed of the input shaft as well as overcoming the drag forces increases the wear on the 1st synchronizer. The wear distribution is 58:42 on the inner and the outer synchronizer rings. The wear of all gear positions is shown in Fig. 11. The gearbox meets the useful life target of 100,000 gear shift cycles in the GSD bench test in each gear. This will be equal to 100,000 km of the useful life of the vehicle. 4.1.3. Vehicle level GSD endurance test On the vehicle GSD test the driveline torsional fluctuations are getting included due to no masking provided to isolate the driveline torsional fluctuations. The wear on the 1st synchronizer ring was 0.65 mm. This wear was due to high inertia load in the 2 to 1 gear shifting with quick gear shifting and the very higher downshift gear speed. Because of this, the work done by the synchronizer ring is higher and which in-turn wears the ring. The wear of all gear positions is shown in Fig. 12. The μ-comp test, transmission level GSD and the vehicle GSD test results show the synchronizer ring generates required friction and have the required capacity to do the synchronization. The wear occurs due to the synchronization load. The inclusion of vehicle torsional fluctuations in the vehicle GSD did not contribute any wear. 4.2. Drivetrain vibrations
Wear Gap (mm)
4.2.1. Flywheel and clutch inertia A user driving pattern is identified on the vehicle to have higher fuel efficiency. The vehicle is mostly driven on the 5th and the 6th gears on the highway with a constant low engine speed in the range of 1000 to 1400 rpm (Figs. 13, 14 and 15). The drive rattle is usually worse at higher torque demands and also worse in direct drive [3]. On the 5th and the 6th gears at 1100 rpm the standard flywheel generates 950 rad/s2 angular acceleration. The standard clutch does not dampen the angular acceleration on the 5th gear and amplifies to 1600 rad/s2. On the 4th gear 1600 rad/s2 angular acceleration is transferred to the input shaft at 1250 rpm. The usage of 4th gear in the said rpm is very less and not considered for analysis. In these conditions the standard clutch disc does not fully dampen the engine torsional fluctuations. The fluctuations are directly transferred to the gearbox (Fig. 16). If the torsional vibration imparted by the internal combustion engine exceeds a certain amplitude, the idle rotating loose parts are induced to vibrate within their functionally defined space [2]. The synchronizer ring is excited by the input shaft vibrations and moves within its clearance and causes impact [8]. In this driving condition, the engaged synchronizer ring does not undergo the vibrations as they are supported by the sleeve and the synchronizer cone surfaces. The non-engaged synchronizer rings such as 1st, 2nd, 3rd and 4th rings are more prone to vibrate as they are not supported or guided by sleeve and gear cone (Figs. 2 and 3). The 1.8 1.6 1.4 1.2 1 0.8 0.6 0.4 0.2 0 1
2
3
4
5
Gear Wear Gap Before Test
Wear Gap A!er Test
Fig. 12. Wear gap in vehicle GSD 435
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% Time Ulizaon
40
20
0 N
1
2
3
4
5
6
Gear
Fig. 13. Customer gear utilization pattern
% Time Utilization
40 30 20 10 0
Engine Speed (rpm)
Fig. 14. % Time utlilization in 5th gear
Fig. 15. % Time utilization in 6th gear
torsional fluctuations create surface fatigue on the hitting zone and in turn wears the synchronizer ring [7]. The ring moves linearly and radially in the package space and hitting with the synchronizer cone with very high frequency. The hitting removes the oil film in between the carbon synchronizer lining and the synchronizer cone surfaces. The hitting removes the carbon particles on the liner surface. The angular accelerations can be reduced by using bigger size flywheel and clutch [3]. To reduce the drivetrain vibration, the clutch disc size was increased to 10%. To accommodate the increased clutch disc diameter, the flywheel diameter also was increased to 10%. The increased flywheel reduced the engine angular acceleration. The 10% diameter increased clutch & flywheel dampened the input shaft torsional vibrations to 1100 rad/s2 for the engine speed range of 1100 to 1400 rpm in 5th gear driving condition as shown in Fig. 16. The gearbox was validated with the standard diameter clutch disc and the 10% diameter increased clutch disc. Both the vehicles were tested at a constant speed of 1100 to 1200 rpm for 10,000 km on the 5th gear. On the 3rd gear synchronizer, the wear got reduced from 0.95 mm to 0.45 mm with the 10% diameter increased clutch when compared with the standard diameter clutch. The reduced angular acceleration reduces the free movement of the rings and the wear was getting reduced in the bigger diameter clutch gearbox (Fig. 17). 436
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Fig. 16. Engine & gearbox angular acceleration
4.2.2. TV test The drivetrain vibrations are simulated using TV test rig. On the TV test set-up, the gearbox with 100 rad/s2 angular acceleration generates 0.05 mm wear on the 3rd gear synchronizer position and no wear in the other gear position, when tested for 65 h. The same gearbox when tested with 2000 rad/s2 angular acceleration for 35 h had a wear of 1.4 mm on the 3rd gear synchronizer position which was in line with the 20,000Km highway durability vehicle gearbox. On, further testing, all the synchronizer rings wear out fully at 65 h as shown in Fig. 18, Fig. 19. Under the magnification of 500× on the microscope, the vehicle level GSD test synchronizer ring wear pattern shows carbon particles with oil filled pores and it is due to pressure loading of synchronizer ring (Fig. 20a). When the vehicle is driven continuously on the 5th gear on the highway, the vehicle failed synchronizer ring does not show any pressure loading. Due to vibration, the synchronizer ring moves freely and hits on the synchronizer cone surface. The hitting on the gear cone causes abrasion wear and removes the top layer of carbon and it gets replaced with the next layer of carbon particles. This reduces the life of the synchronizer ring. The new synchronizer, the highway durability vehicle failed synchronizer, and the TV test rig synchronizer have the same 437
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1
Wear (mm)
0.8 0.6 0.4 0.2 0 1
2
3
4
5
6
Gear Standard Ø Clutch
10% Increased Ø Clutch
Fig. 17. Influence of clutch size on wear reduction
Fig. 18. TV simulation comparison with vehicle test
Fig. 19. a) Nominal wear gap, b) Zero wear gap
Fig. 20. a) GSD Test completed carbon surface–loaded carbon surface, b) Torsional vibration failed non-loaded carbon surface
carbon pattern (Fig. 20b). 4.2.3. Center support of synchronizer ring In the dual cone synchronizer ring, the outer cone is guided by the sleeve and or hub with a nominal radial clearance of 0.5 mm. The inner cone has a radial clearance of 3 mm with the surrounding parts such as hub or gear. Due to this the inner cone can move freely in the radial direction. On the high way durability test, the wear of the inner carbon liner contributes to 85% of total synchronizer wear. The inner cone radial clearance with the hub is reduced to 0.5 mm as shown in Fig. 21. This reduces the free movement of the inner synchronizer ring and gives good guidance to the inner synchronizer ring. On the TV test rig for 65 h, the 438
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Fig. 21. Center support for inner synchronizer ring
gearbox with 0.5 mm radial clearance had a wear of 0.98 mm compared with 1.4 mm wear of 3 mm radial clearance synchronizer ring (Fig. 22). Due to center support the wear of the inner and the outer carbon liners became 60:40 ratio. The reduced clearances reduced the radial movement of the synchronizer rings and reduced the intensity of the hitting of the synchronizer ring and in turn reduced the wear in all gear synchronizer positions. 4.2.4. Oil viscosity Kinematic viscosity has a major influence on the loose part vibration generated by the transmissions. Correct selection of lubricant enables the loose part vibration limit to be moved to higher angular acceleration amplitudes and vibration can be reduced [2]. On the TV test rig, the gearbox was tested with SAE80 oil and SAE80W90 oil for 65 h. The oil parameters are given in Table 9. For the low viscosity, the movement of the ring is higher due to less dampening in the synchronizers contact [8]. The high viscosity oil has higher compressibility which results in higher load carrying capacities. The thicker oil provides an oil cushion for contact [9]. The oil viscosity increases the drag in the gearbox and reduces the movement of loose parts [3]. In 65 h of TV test, the SAE80 oil had a wear of 0.98 mm compared with 0.52 mm wear in SAE80W90 oil for the 3rd synchronizer position. The wear got reduced in SAE80W90 oil compared with SAE80 oil in all gear synchronizer positions as shown in Fig. 23. 4.2.5. Oil volume Due to torsional vibration the intermediate synchronizer ring vibrates within the available space and causes hammering effect. The oil film in between the carbon liner and the cone breaks away and leads to wear of carbon liner. The synchronizer ring can move within the available package space. In the gearbox the synchronizer ring can move an axial distance of 1 mm. This axial movement creates a radial gap of 0.15 mm. The outer ring can move 0.5 mm inside the hub and the inner ring can move 0.5 mm towards the hub center support. Sufficient oil quantity should be available in the gaps (Fig. 24) to dampen the vibrations. The oil level in the gearbox is to be optimized to have enough oil movement within the gearbox during churning lubrication [10]. The level of lubrication oil in the gearbox is of importance, as it develops drag torque thus reducing the vibration of loose parts [2,11]. The drag torque is generated by the shear torque of lubrication in the gap (or gaps) between the synchronizer ring and the synchronizer cone, which are rotating at different speeds. The axial velocity of the oil due to the centrifugal force is given as below [12].
u (Z ) =
k1 k 0 Z (0.5
2
+2
2 a
+2
a
(8)
)
(9)
Wear (mm)
k1 = tan
1.6 1.4 1.2 1 0.8 0.6 0.4 0.2 0 1
2
3
4
Gear Clearance 0.5mm, SAE80, 5L Oil
Clearance 3mm, SAE80, 5L Oil
Fig. 22. Influence of center support on synchronizer wear 439
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Table 9 Oil properties Parameter
SAE80
SAE 80W90
Kinematic viscosity, mm2/s at 40 °C Kinematic viscosity, mm2/s at 100 °C Base oil density kg/dm3
74.6 10.2 0.85
135 16 0.88
1.2
Wear (mm)
1 0.8 0.6 0.4 0.2 0 1
2
Clearance 0.5mm, SAE80, 5L Oil
3
4
Gear Clearance 0.5mm, SAE80W90, 5L Oil
Fig. 23. Influence of oil viscosity on synchronizer wear
Fig. 24. Synchronizer in neutral position with oil flow
k 0 = rm +
1 bsin 2
(10)
Axial volumetric flow rate due to the centrifugal force is as follows [12].
Q = 2 h ( k1 Z + k 0
(11)
0.5h) u (Z )
The axial volumetric flow increases linearly by increasing the space between the synchronizer ring to have fully wetted surface. When the oil full film wetting surface of synchronizer ring width increases, the axial volumetric flow needed to maintain full film wetted volume also increases as shown in Fig. 25. The full film wetted surface area generates higher drag than the oil-air mixture in the annular space of synchronizer ring. To increase the full film wetted surface area the oil volume and or the oil level in the gearbox is to be increased. A CFD analysis of gearbox was performed on the 3rd and the 4th gear positions. The purpose was to increase the drag (reduce the torsional vibration) and improve lubrication/oil flow in the synchronizer components. Churning loss is a speed dependent loss that will produce higher gearbox drag at higher speeds as large quantity of oil has to be moved per time unit [11]. The oil resists movement of parts, there by dampening vibrations and mechanical shocks. The oil volume in the gearbox was 5 l. The volume of the oil was increased to 7 l to provide higher splash in to wear gap, better lubrication and to dampen the vibrations. The 5 l CFD analysis showed less oil movement around the 3rd and the 4th synchronizer rings. The 7 l CFD analysis revealed that more oil reached the friction lining material on the 4th gear side via the openings of clutch body ring lugs. The opening of clutch body ring lugs in the 3rd gear side was covered by the larger 3rd gear, so less quantity of oil was reaching the lining material as shown in Fig. 26, Fig. 27. The 4th counter shaft gear had larger diameter and had higher immersion depth on the oil compared with the 3rd gear. It drew up more oil and created more splash and stirred the oil bath more [13]. In the TV 440
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Q (L/min)
B. K, et al.
0.9 0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1 0 0
0.02
0.04
0.06
0.08
0.1
0.12
0.14
0.16
h (mm) z=4
z=5
z=6
z=7
Fig. 25. Volume of Oil required for full wetting of gear cone surface for different gear cone to synchronizer ring gaps
Fig. 26. Oil flow distribution on the 3rd & the 4th gear positions for 5 l and 7 l of oil (section taken between synchronizer ring and CB ring).
test for 65 h, the 3rd synchronizer had a wear of 0.52 mm in 5 l gearbox and the wear reduced to 0.3 mm in 7 l gearbox. For the 4th gear the wear was 0.5 mm in the 3rd gear and the wear reduced to 0.2 mm in the 7 l oil gearbox as shown in Fig. 28. This wear reduction happened due to the higher oil level and higher oil churning. The increased level of oil volume increased the drag of counter shaft gears and increased the drag torque on synchronizers [14]. This reduced the free movement of the synchronizer rings by dampening the vibration transferring from the 5th gear to counter shaft gears and in turn reduced the impact in all gear synchronizer positions. 4.2.6. Engine dyno test The improvements carried out in the study such as clutch diameter increase, reduced inner synchronizer ring clearance, oil grade and oil volume were validated in the engine dyno test. The initial design had a wear of 0.9 mm in the 3rd synchronizer position on the engine dyno. This wear reduced to 0.1 mm with 441
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Fig. 27. Oil flow distribution on the 3rd & the 4th gear positions
0.6
Wear (mm)
0.5 0.4 0.3 0.2 0.1 0 1
2
3
4
Gear Clearance 0.5mm, SAE80W90, 5L Oil
Clearance 0.5mm, SAE80W90, 7L Oil
Fig. 28. Influence of oil volume on synchronizer wear
the 10% increased clutch diameter, 0.5 mm inner synchronizer ring clearance, with 7 l of SAE80W90 oil as shown in Fig. 29. The wear was reduced in all gear synchronizer positions. 5. Conclusion This paper describes the synchronizer ring wear due to torsional vibration in manual transmission gearbox synchronizers. The experimental study reduces the synchronizer wear. The conclusions from the study and the experimental results are as follows:
1
Wear (mm)
0.8 0.6 0.4 0.2 0 1
2
3
4 5 6 Gear Std. Dia Clutch, Clearance 3mm, SAE80, 5 Ltr oil 10% increaesd Dia Clutch, Clearance 0.5mm, SAE80W90, 7 Ltr oil Fig. 29. Configuration of engine dyno test and synchronizer wear 442
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1. The torsional vibration causes non-uniform wear on the synchronizer ring. The inner ring wear is higher than the outer ring wear. 2. The transfer of torsional vibration into the gearbox is dampened by the bigger diameter clutch. The inner synchronizer ring guide restricts the free movement of the ring. Further high viscosity oil dampens the vibration which cause impact of synchronizer ring and wear. 3. More amount of oil is getting churned with the increased oil volume, which pushes more oil in to the synchronizer ring and dampen the vibration. The increase in the drag on the counter shaft gears further cut the vibration transfer path. 6. Future Scope of work Increase of clutch size increases the downshift reflected inertia. Oil volume and the oil viscosity increases the drag of the gearbox. These will affect the gear shift quality, which can be solved by optimizing the synchronizer index percentage, tooth thickness and tooth chamfer angles. Nomenclature A b dmax Fs Grpm Hrpm h rm IRef Nc Q u(z) Z α ∆rpm ω ωa ωi
Surface area (mm2) width of synchronizer ring (mm) Max cone diameter (mm) Force at sleeve (N) Gear speed (rpm) Hub speed (rpm) radial space between the gear cone and synchronizer ring (mm) Mean radius of synchronizer ring (mm) Reflected inertia of the gear (Kgm2) Number of cones Axial volumetric flow rate (mm3) Axial velocity (mm/s2) Full oil film length (mm) Cone Angle (°) Difference in speed (rpm) Angular acceleration (rad/s2) Angular acceleration of the synchronizer ring / hub (rad/s2) Angular acceleration of the gear (rad/s2)
References [1] Paul D. Walker and Nong Zhang, “Transmission of engine harmonics to synchronizer mechanisms in dual clutch transmissions”, J. Vib. Acoust., Paper No: VIB14-1015; doi: https://doi.org/10.1115/1.4028079. [2] S.N. Dogan, J. Ryborz, G. Lechner, The Effect of Transmission Lubricants on Rattling and Clattering Noise in Automotive Transmissions, WTC World Tribology Congress, Vienna, Austria, 2001 (03.-07.09.01). [3] R. Seaman, C. Johnson, R. Hamilton, Component inertial effects on transmission design, SAE Tech. Pap. (1984), https://doi.org/10.4271/841686 841686. [4] W.R. Reik, The dual mass flywheel, 6th Luk Symposium, 1998. [5] N. Tawhare, D. Aswale, M. Ponkshe, Torsional fluctuations consideration while design of synchro rings, SAE Tech. Pap. (2018), https://doi.org/10.4271/201801-1823 2018-01-1823. [6] R. Socin, L. Walters, Manual transmission synchronizers, SAE Tech. Pap. (1968), https://doi.org/10.4271/680008 680008. [7] Singh, M., Singh, P., Kumar, K., and Rawla, V., "Development of test method to validate Synchronizer Ring design for torsional fluctuations in manual transmission," SAE Tech. Pap. (2016-28-0012). [8] P. Fietkau, A. Baumann, B. Bertsche, Simulation of passenger car synchronizer ring movement during rattling, Proc. Inst. Mech. Eng. Part K 226 (2012) 3–16, https://doi.org/10.1177/1464419311428394. [9] Erwin Adi Hartono, Study of Fluid Flow Inside the Gearbox, PhD Thesis Department of Applied Mechanics Division of Fluid Dynamics Chalmers University Of Technology, Sweden, 2014. [10] Hua Liu, Thomas Jurkschat, Thomas Lohner, Karsten Stahl, Detailed investigations on the oil flow in dip-lubricated gearboxes by the finite volume CFD method, Lubricants 6 (2) (2018) 47, https://doi.org/10.3390/lubricants6020047. [11] Keerthi Medleri Hire Math, Mikael Lund, Drag Torque and Synchronization Modelling in a Dual Clutch Transmission, Master Thesis, Chalmers University Of Technology & Lund University. [12] Y. Shen, Z. Liu, S. Rinderknecht, Modelling of power losses of transmission synchronizers in neutral position, SAE Tech. Pap. (2018), https://doi.org/10.4271/ 2018-01-1228 2018-01-1228. [13] E.A. Hartono, M. Golubev, V. Chernoray, PIV study of fluid flow inside a gearbox, Proc. of 10th International Symposium on Particle Image Velocimetry, 2013, pp. 1–11. [14] Z. Liu, Y. Shen, S. Rinderknecht, Theoretical and experimental investigation on power loss of vehicle transmission synchronizers with spray lubrication, SAE Tech. Pap. (2019), https://doi.org/10.4271/2019-01-0028 2019-01-0028.
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Engineering Failure Analysis journal homepage: www.elsevier.com/locate/engfailanal
Analysis of automotive transmission gearbox synchronizer wear due to torsional vibration and the parameters influencing wear reduction
T
Barathiraja Ka,b, , Devaradjane Gb, Jibin Paula, Rakesh Sa, Gajanan Jamadadea ⁎
a b
Mahindra & Mahindra Ltd, Mahindra Research Valley, Chennai, India Department of Automobile Engineering, Madras Institute of Technology, Anna University, Chennai, India
ARTICLE INFO
ABSTRACT
Keywords: Manual transmission Carbon synchronizers Wear Torsional vibration Lubrication Angular acceleration
Synchronizers are the heart for the manual transmission, automated manual transmission and dual clutch transmission. The synchronizers match the speed of the target gears during gear shifts. Downsizing the high-power density engine develops higher angular accelerations. Higher angular accelerations create torsional vibrations and are detrimental to the life of the synchronizers. The synchronizer rings can move freely in the available space due to torsional vibration. The synchronizers which experience higher angular acceleration collide with the surrounding parts and wear out. The wear of synchronizer carbon liner reduces the wear gap to zero. The zerowear gap hampers the synchronizer functionality and leads to gear clash. This paper presents the impact of angular acceleration on the life of carbon synchronizer ring and the parameters which are influencing to overcome the failure are studied. A bench test set-up was developed to simulate the vehicle level angular accelerations. The bench test results show direct correlation with the life of the synchronizer on the vehicle. Torsional vibration dampening using clutch size, oil viscosity, guiding of synchronizer ring, and oil volume are studied. The synchronizer carbon liner wear reduction is studied and validated with different bench and vehicle level tests.
1. Introduction Synchronizers are conical mechanical clutches used to synchronize the speed of the rotating parts. Synchronizers are used in manual transmission, automated manual transmission, and dual clutch transmissions for passenger and commercial vehicles. The most commonly used material for synchronizer is brass. To improve the friction of cone surface, wear resistance, and durability of synchronizer, carbon synchronizers are widely used. Under severe test and operating conditions, the carbon synchronizer shows higher wear resistance and durability when compared with brass synchronizers. In vehicles, the internal combustion engines are the main source for the angular acceleration and torsional fluctuations due to the engine firing order [1,2]. Stringent emission norms and the demand for fuel efficient engines force the automotive industries to downsize the engines. Small sized high-power density internal combustion engines generate high torsional vibration [3]. The torsional fluctuation generated by the engine is transmitted to the gearbox through clutch. Dual mass flywheel can filter the torsional fluctuations produced by the engine [4]. However, generally not used in commercial vehicle engines, due to high cost when
Abbreviations: TV, Torsional vibration; GSD, Gear shift durability.; CFD, Computational fluid dynamics ⁎ Corresponding author at: Mahindra & Mahindra Ltd, Mahindra Research Valley, Chennai, India. E-mail address: [email protected] (B. K). https://doi.org/10.1016/j.engfailanal.2019.06.084 Received 8 March 2019; Received in revised form 25 June 2019; Accepted 25 June 2019 Available online 26 June 2019 1350-6307/ © 2019 Elsevier Ltd. All rights reserved.
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Fig. 1. 6 Speed gearbox layout
compared with single mass flywheel. All the fixed gearwheels in the commercial vehicle coaxial manual transmission gearbox are mounted on the countershaft. This leads to excitation of all the idler gears by the countershaft, both in neutral and with any gear step engaged condition [1]. The torsional vibration creates the brass synchronizer ring lug failure and synchronizer ring wear gap reduction [5,7]. This paper reports on an experimental investigation of carbon synchronizer liner wear reduction in a conventional manual gearbox under torsional vibration condition. First, the endurance capacity or amount of wear on the designed synchronizer are studied with the μ-comp test rig, bench level GSD test rig and vehicle GSD test. Then the vehicle is instrumented to capture the driver driving pattern such as gear usage pattern and the engine speed. The gearbox and the engine are instrumented to capture the torsional vibrations on the engine flywheel and input shaft of the gearbox after the clutch. This measurement describes the input torsional fluctuations to the gearbox. Then the gearbox is instrumented and tested on the TV test rig which replicates the actual vehicle carbon synchronizer liner wear. On the TV test rig, the gearbox is tested on different clutch sizes, oil volume, oil viscosity and inner diameter guiding of synchronizer ring. The improvements are validated on the vehicle to confirm meeting of 100,000 km target distance. 2. Statement of problem The gearbox taken for the study is inline manual transmission having input shaft, counter shaft and output shaft. It has dual cone carbon synchronizers in the 1st, 2nd, 3rd and 4th gear positions and single cone carbon synchronizer in 5th and 6th gear positions. All the synchronizers are located on the output shaft as shown in Fig. 1. The synchronizer assembly consists of synchronizer hub, synchronizer sleeve, synchronizer blocker ring, dog teeth and strut detent. The outer synchronizer ring and inner synchronizer ring are made of steel and have cone surface. The carbon liner is bonded on the steel intermediate synchronizer ring. The outer, intermediate and inner synchronizer ring parts are assembled loosely with in the available space of hub to gear assembly as shown in Figs. 2 and 3. The synchronizer assembly is designed to have a nominal wear gap of 1.5 mm. The useful life target of the gearbox is 100,000 km. The synchronizer rings wear off quickly in the vehicle validation in less than 20,000 km especially in 1st, 2nd, 3rd & 4th gear positions. Due to ring wear, the wear gap as shown in Fig. 3, is getting closed or become zero. Zero wear gap hampers synchronization and the gearbox results into early gear crash on the non-engaged synchronizer rings. 3. Experimental set-up The experimental test flow is given in Fig. 4. The synchronizer design is validated for the coefficient of friction generation and the endurance life with the help of μ-comp synchronizer endurance test, Transmission level GSD test and Vehicle level GSD. These tests do not include any external noise factors. Then the external factors which affects the synchronizer function are checked using customer driving pattern analysis, TV test and Highway durability test. The above experimental test flow can be broadly divided into two types: Bench level test and Vehicle level test according to Table 1. The bench tests are the replica of the vehicle test. The bench test gives the flexibility of controlling the test parameters and reduces the total test duration. The external noise factors are not present in the bench test. However, the vehicle tests are carried out to simulate the actual vehicle driving conditions.
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Fig. 2. Synchronizer assembly
Fig. 3. Synchronizer ring wear gap
3.1. Bench tests 3.1.1. μ-comp Synchronizer endurance test rig The component level endurance test is conducted using a conventional μ-comp test rig as shown in Fig. 5. The synchronizer hub is mounted on the main shaft of the rig. The main shaft is connected to the electric motor. The sleeve slides over the hub. A hydraulic shift actuator is used to move the sleeve. The sleeve axial force and the linear travel are measured using force sensor and linear accelerometers. The outer and the inner synchronizer rings are connected to the inertia disc through the hub assembly. The intermediate synchronizer ring is connected to the gear and it is rigidly fixed to the test rig. An RPM sensor measures the main shaft speed. The lubrication oil is supplied over the synchronizer ring through a pump. The lubrication oil flow rate is measured and controlled using a flow meter. A heater is provided in the hydraulic system to heat up the oil during test. The cone torque generated on the synchronizer cone surfaces is measured using a strain gauge. The coefficient of friction and the wear are calculated by the system. The test specification is according to Table 2. 3.1.2. Transmission GSD test rig On the GSD test rig (Fig. 6), the electric motor is coupled to the output shaft. The clutch disc is mounted on the input shaft. Two 429
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Fig. 4. Block diagram of test flow sequence Table 1 Bench tests and vehicle tests Bench tests
Vehicle tests
μ-comp synchronizer endurance test Transmission GSD Test TV test rig Engine Dyno test
Vehicle GSD test Driver driving pattern measurement Vehicle angular acceleration measurement High way durability test
Fig. 5. μ-comp test rig setup
pneumatic actuators are connected to the gear shift knob for the gear select and gear shift directions. The input and the output shaft rpm are measured using speed sensors. The gear shift forces and the travel are measured using strain gauges and position sensors. A temperature sensor is used to monitor the transmission oil temperature. For dynamic gearshift, the speed of the input shaft during gear shift is as per the Table 3. The gears are shifted in the sequence according to Table 4. 430
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Table 2 Test specification for μ-comp test Parameter
Specification
Mean cone diameter Cone angle Force Delta speed Inertia Cycle time Test cycle
110 mm 8.5° 2000 N 1500 rpm 0.125 Kgm2 3.0 s 100,000
Fig. 6. Gear shift durability test set-up Table 3 GSD test bench gear shift speeds Shift type
Input shaft speed (rpm)
Up-shift Down-shift
1800 1200
Table 4 Gearbox GSD test cycle Gear shift sequence
No of cycles
1-2-1 2-3-2 3-4-3 4-5-4 5-6-5
50,000 100,000 100,000 100,000 100,000
3.1.3. TV test rig The TV (Torsional Vibration) test rig consists of gearbox, drive motor, propeller shaft & adopter pivot as shown in Fig. 7. The clutch disc is mounted on the gearbox input shaft. The electric motor is connected to the gearbox output shaft through the coupling and the propeller shaft. The adopter plate to the mount is offset by 15°-20° on the vertical axis and 10°-15° in the horizontal axis. The propeller shaft yokes have a phase difference of 90° as shown in Fig. 8. This gives a phase difference during shaft rotation. A magnetic pickup speed sensor is used to measure the input shaft speed. The acceleration is processed using sensor electronics. The gearbox is engaged in the 5th gear position (1:1 gear ratio). The motor is continuously driven at 1100 rpm. This rpm with propeller shaft phase difference creates 2000 rad/s2 angular acceleration on the input shaft. This set-up vibrates the synchronizer ring assembly. The synchronizer life in the vehicle can be directly correlated with the TV test rig. A DOE is performed using one factor at a time DOE methodology. The test sequence on the TV test rig is as per the Table 5. 431
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Fig. 7. Torsional vibration test set-up
Fig. 8. Propeller shaft phase out Table 5 Test configuration of TV test rig Test No
Hub to ring clearance (mm)
Oil grade
Oil volume (l)
1 2 3 4
3 0.5 0.5 0.5
SAE80 SAE80 SAE80W90 SAE80W90
5 5 5 7
3.1.4. Engine Dyno test The test transmission is coupled to the engine on the engine dyno test bed as shown in Fig. 9. The gearbox is pre-selected to the 5th gear position. The engine power is transferred to the gearbox through the flywheel & clutch. The engine rpm is measured with a speed sensor on the flywheel. The engine is continuously running at 1100–1200 rpm for 300 h. The test configuration for the gearbox is as per the Table 6. 3.2. Vehicle test The gearbox is mounted on a vehicle to test the gearbox in the actual driving conditions. The vehicle is loaded to the GVW to simulate the actual load condition. 3.2.1. Vehicle GSD test In the vehicle the gears are shifted in the sequence of 1-2-3-4-5-6-5-4-3-2-1. In this test, trained drivers do quick gear shift at an 432
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Fig. 9. Engine dyno test set-up Table 6 Test configuration of engine dyno Test no
Clutch size
Hub to ring clearance (mm)
Oil grade
Oil volume (l)
1 2
Standard diameter 10% increased diameter
3 0.5
SAE80 SAE80W90
5 7
engine speed as per the Table 7. The vehicle was driven for 15,000 shift cycles. This is an abuse test for the synchronizer components. 3.2.2. Vehicle angular acceleration On the vehicle, the engine and the gearbox are instrumented to measure the angular accelerations. A magnetic pickup tachometer is mounted on the gearbox focusing the engine flywheel tooth. Another magnetic pickup tachometer was positioned in the gearbox focusing the input shaft gear. This is used to measure the speed, in revolutions per minute. Whenever the tachometer sensor detects a gear tooth, it generates a pulse. The tachometers are connected to SCADAS, and LMS Test Lab is used to measure the angular accelerations of the flywheel and the input shaft gear. 3.2.3. Vehicle high way durability test The vehicle is driven on the high-speed durability test cycle. Trained drivers are driving the vehicle on the highway. The vehicle is quickly taken into 5th gear by quick gear shifting from “N” gear. Then the vehicle is driven at an engine speed of 1100–1400 rpm on 5th gear. 4. Results & Discussion The synchronizer ring wear can occur due to poor design and or due to external factors. The synchronizer capacity is used to validate the robustness in design. The external factors which causes wear are discussed under drivetrain vibrations. 4.1. Synchronizer capacity The synchronizer capacity was validated to check the life of the synchronizer ring. This validates the capability of the rings to Table 7 Vehicle GSD gear shift speeds Shift type
Input shaft speed (rpm)
Up-shift Down-shift
1900 1900
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Table 8 Material limits of carbon synchronizer Specification
Material Limits
Energy limit Surface pressure Surface velocity
1.5 J/mm2 10 N/mm2 8 to 10 m/s
perform its intended function of synchronization. 4.1.1. Synchronizer level GSD test The synchronizer ring assembly was validated in μ-comp test rig for the endurance life. To meet the endurance life of the gearbox, the material limits of the synchronizer for each gear shift operations should be more than the synchronizer capacity as per the Table 8. The synchronizer capacity material limits are calculated using the expression [6].
Energy Limit =
1 IRef 2
Surface Pressure =
2
Fs Nc A sin
Surface Velocity = dmax
= abs ( rpm
i
= abs (Grpm
(1) (2) rpm
(3)
60
(4)
a)
(5)
Hrpm )
This test was performed in a controlled environment with no vibration in the system. The test gives a stable friction and μdynamic is 0.105 at the end stage. The total wear is 0.21 mm for the entire test cycle (Fig. 10). The wear distribution is 55:45 on the inner and outer rings with no abnormalities in the entire test cycle. When the vehicle is driven on top gear with max engine rpm, the idle synchronizer packs in the low gear experience a high surface velocity. The 1st gear experiences 30% higher surface velocity than the carbon limits. In the next test, the hub speed is increased to match the vehicle Δrpm on the test rig without performing the gear shift operation. This generates 30% higher surface velocity on the synchronizer cone surfaces. The test was conducted for 5 h and no wear observed. 4.1.2. Transmission level GSD Endurance test The wear of 1st synchronizer ring in the 1-2-1 gear shift is 0.55 mm. The 1st gear reflected inertia is higher than other gears. The
Fig. 10. μ-comp coefficient of friction 434
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Wear Gap (mm)
B. K, et al.
1.8 1.6 1.4 1.2 1 0.8 0.6 0.4 0.2 0 1
2
3
4
5
6
Gear Wear Gap Before Test
Wear Gap A!er Test
Fig. 11. Wear gap in transmission GSD
higher inertia causes higher wear on the synchronizer rings when compared to other gear positions. During upshift the gearbox drag helps for synchronization. In downshift the gearbox has to work against the drag for synchronization. The higher work done by the synchronizer by accelerating the speed of the input shaft as well as overcoming the drag forces increases the wear on the 1st synchronizer. The wear distribution is 58:42 on the inner and the outer synchronizer rings. The wear of all gear positions is shown in Fig. 11. The gearbox meets the useful life target of 100,000 gear shift cycles in the GSD bench test in each gear. This will be equal to 100,000 km of the useful life of the vehicle. 4.1.3. Vehicle level GSD endurance test On the vehicle GSD test the driveline torsional fluctuations are getting included due to no masking provided to isolate the driveline torsional fluctuations. The wear on the 1st synchronizer ring was 0.65 mm. This wear was due to high inertia load in the 2 to 1 gear shifting with quick gear shifting and the very higher downshift gear speed. Because of this, the work done by the synchronizer ring is higher and which in-turn wears the ring. The wear of all gear positions is shown in Fig. 12. The μ-comp test, transmission level GSD and the vehicle GSD test results show the synchronizer ring generates required friction and have the required capacity to do the synchronization. The wear occurs due to the synchronization load. The inclusion of vehicle torsional fluctuations in the vehicle GSD did not contribute any wear. 4.2. Drivetrain vibrations
Wear Gap (mm)
4.2.1. Flywheel and clutch inertia A user driving pattern is identified on the vehicle to have higher fuel efficiency. The vehicle is mostly driven on the 5th and the 6th gears on the highway with a constant low engine speed in the range of 1000 to 1400 rpm (Figs. 13, 14 and 15). The drive rattle is usually worse at higher torque demands and also worse in direct drive [3]. On the 5th and the 6th gears at 1100 rpm the standard flywheel generates 950 rad/s2 angular acceleration. The standard clutch does not dampen the angular acceleration on the 5th gear and amplifies to 1600 rad/s2. On the 4th gear 1600 rad/s2 angular acceleration is transferred to the input shaft at 1250 rpm. The usage of 4th gear in the said rpm is very less and not considered for analysis. In these conditions the standard clutch disc does not fully dampen the engine torsional fluctuations. The fluctuations are directly transferred to the gearbox (Fig. 16). If the torsional vibration imparted by the internal combustion engine exceeds a certain amplitude, the idle rotating loose parts are induced to vibrate within their functionally defined space [2]. The synchronizer ring is excited by the input shaft vibrations and moves within its clearance and causes impact [8]. In this driving condition, the engaged synchronizer ring does not undergo the vibrations as they are supported by the sleeve and the synchronizer cone surfaces. The non-engaged synchronizer rings such as 1st, 2nd, 3rd and 4th rings are more prone to vibrate as they are not supported or guided by sleeve and gear cone (Figs. 2 and 3). The 1.8 1.6 1.4 1.2 1 0.8 0.6 0.4 0.2 0 1
2
3
4
5
Gear Wear Gap Before Test
Wear Gap A!er Test
Fig. 12. Wear gap in vehicle GSD 435
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% Time Ulizaon
40
20
0 N
1
2
3
4
5
6
Gear
Fig. 13. Customer gear utilization pattern
% Time Utilization
40 30 20 10 0
Engine Speed (rpm)
Fig. 14. % Time utlilization in 5th gear
Fig. 15. % Time utilization in 6th gear
torsional fluctuations create surface fatigue on the hitting zone and in turn wears the synchronizer ring [7]. The ring moves linearly and radially in the package space and hitting with the synchronizer cone with very high frequency. The hitting removes the oil film in between the carbon synchronizer lining and the synchronizer cone surfaces. The hitting removes the carbon particles on the liner surface. The angular accelerations can be reduced by using bigger size flywheel and clutch [3]. To reduce the drivetrain vibration, the clutch disc size was increased to 10%. To accommodate the increased clutch disc diameter, the flywheel diameter also was increased to 10%. The increased flywheel reduced the engine angular acceleration. The 10% diameter increased clutch & flywheel dampened the input shaft torsional vibrations to 1100 rad/s2 for the engine speed range of 1100 to 1400 rpm in 5th gear driving condition as shown in Fig. 16. The gearbox was validated with the standard diameter clutch disc and the 10% diameter increased clutch disc. Both the vehicles were tested at a constant speed of 1100 to 1200 rpm for 10,000 km on the 5th gear. On the 3rd gear synchronizer, the wear got reduced from 0.95 mm to 0.45 mm with the 10% diameter increased clutch when compared with the standard diameter clutch. The reduced angular acceleration reduces the free movement of the rings and the wear was getting reduced in the bigger diameter clutch gearbox (Fig. 17). 436
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Fig. 16. Engine & gearbox angular acceleration
4.2.2. TV test The drivetrain vibrations are simulated using TV test rig. On the TV test set-up, the gearbox with 100 rad/s2 angular acceleration generates 0.05 mm wear on the 3rd gear synchronizer position and no wear in the other gear position, when tested for 65 h. The same gearbox when tested with 2000 rad/s2 angular acceleration for 35 h had a wear of 1.4 mm on the 3rd gear synchronizer position which was in line with the 20,000Km highway durability vehicle gearbox. On, further testing, all the synchronizer rings wear out fully at 65 h as shown in Fig. 18, Fig. 19. Under the magnification of 500× on the microscope, the vehicle level GSD test synchronizer ring wear pattern shows carbon particles with oil filled pores and it is due to pressure loading of synchronizer ring (Fig. 20a). When the vehicle is driven continuously on the 5th gear on the highway, the vehicle failed synchronizer ring does not show any pressure loading. Due to vibration, the synchronizer ring moves freely and hits on the synchronizer cone surface. The hitting on the gear cone causes abrasion wear and removes the top layer of carbon and it gets replaced with the next layer of carbon particles. This reduces the life of the synchronizer ring. The new synchronizer, the highway durability vehicle failed synchronizer, and the TV test rig synchronizer have the same 437
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1
Wear (mm)
0.8 0.6 0.4 0.2 0 1
2
3
4
5
6
Gear Standard Ø Clutch
10% Increased Ø Clutch
Fig. 17. Influence of clutch size on wear reduction
Fig. 18. TV simulation comparison with vehicle test
Fig. 19. a) Nominal wear gap, b) Zero wear gap
Fig. 20. a) GSD Test completed carbon surface–loaded carbon surface, b) Torsional vibration failed non-loaded carbon surface
carbon pattern (Fig. 20b). 4.2.3. Center support of synchronizer ring In the dual cone synchronizer ring, the outer cone is guided by the sleeve and or hub with a nominal radial clearance of 0.5 mm. The inner cone has a radial clearance of 3 mm with the surrounding parts such as hub or gear. Due to this the inner cone can move freely in the radial direction. On the high way durability test, the wear of the inner carbon liner contributes to 85% of total synchronizer wear. The inner cone radial clearance with the hub is reduced to 0.5 mm as shown in Fig. 21. This reduces the free movement of the inner synchronizer ring and gives good guidance to the inner synchronizer ring. On the TV test rig for 65 h, the 438
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Fig. 21. Center support for inner synchronizer ring
gearbox with 0.5 mm radial clearance had a wear of 0.98 mm compared with 1.4 mm wear of 3 mm radial clearance synchronizer ring (Fig. 22). Due to center support the wear of the inner and the outer carbon liners became 60:40 ratio. The reduced clearances reduced the radial movement of the synchronizer rings and reduced the intensity of the hitting of the synchronizer ring and in turn reduced the wear in all gear synchronizer positions. 4.2.4. Oil viscosity Kinematic viscosity has a major influence on the loose part vibration generated by the transmissions. Correct selection of lubricant enables the loose part vibration limit to be moved to higher angular acceleration amplitudes and vibration can be reduced [2]. On the TV test rig, the gearbox was tested with SAE80 oil and SAE80W90 oil for 65 h. The oil parameters are given in Table 9. For the low viscosity, the movement of the ring is higher due to less dampening in the synchronizers contact [8]. The high viscosity oil has higher compressibility which results in higher load carrying capacities. The thicker oil provides an oil cushion for contact [9]. The oil viscosity increases the drag in the gearbox and reduces the movement of loose parts [3]. In 65 h of TV test, the SAE80 oil had a wear of 0.98 mm compared with 0.52 mm wear in SAE80W90 oil for the 3rd synchronizer position. The wear got reduced in SAE80W90 oil compared with SAE80 oil in all gear synchronizer positions as shown in Fig. 23. 4.2.5. Oil volume Due to torsional vibration the intermediate synchronizer ring vibrates within the available space and causes hammering effect. The oil film in between the carbon liner and the cone breaks away and leads to wear of carbon liner. The synchronizer ring can move within the available package space. In the gearbox the synchronizer ring can move an axial distance of 1 mm. This axial movement creates a radial gap of 0.15 mm. The outer ring can move 0.5 mm inside the hub and the inner ring can move 0.5 mm towards the hub center support. Sufficient oil quantity should be available in the gaps (Fig. 24) to dampen the vibrations. The oil level in the gearbox is to be optimized to have enough oil movement within the gearbox during churning lubrication [10]. The level of lubrication oil in the gearbox is of importance, as it develops drag torque thus reducing the vibration of loose parts [2,11]. The drag torque is generated by the shear torque of lubrication in the gap (or gaps) between the synchronizer ring and the synchronizer cone, which are rotating at different speeds. The axial velocity of the oil due to the centrifugal force is given as below [12].
u (Z ) =
k1 k 0 Z (0.5
2
+2
2 a
+2
a
(8)
)
(9)
Wear (mm)
k1 = tan
1.6 1.4 1.2 1 0.8 0.6 0.4 0.2 0 1
2
3
4
Gear Clearance 0.5mm, SAE80, 5L Oil
Clearance 3mm, SAE80, 5L Oil
Fig. 22. Influence of center support on synchronizer wear 439
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Table 9 Oil properties Parameter
SAE80
SAE 80W90
Kinematic viscosity, mm2/s at 40 °C Kinematic viscosity, mm2/s at 100 °C Base oil density kg/dm3
74.6 10.2 0.85
135 16 0.88
1.2
Wear (mm)
1 0.8 0.6 0.4 0.2 0 1
2
Clearance 0.5mm, SAE80, 5L Oil
3
4
Gear Clearance 0.5mm, SAE80W90, 5L Oil
Fig. 23. Influence of oil viscosity on synchronizer wear
Fig. 24. Synchronizer in neutral position with oil flow
k 0 = rm +
1 bsin 2
(10)
Axial volumetric flow rate due to the centrifugal force is as follows [12].
Q = 2 h ( k1 Z + k 0
(11)
0.5h) u (Z )
The axial volumetric flow increases linearly by increasing the space between the synchronizer ring to have fully wetted surface. When the oil full film wetting surface of synchronizer ring width increases, the axial volumetric flow needed to maintain full film wetted volume also increases as shown in Fig. 25. The full film wetted surface area generates higher drag than the oil-air mixture in the annular space of synchronizer ring. To increase the full film wetted surface area the oil volume and or the oil level in the gearbox is to be increased. A CFD analysis of gearbox was performed on the 3rd and the 4th gear positions. The purpose was to increase the drag (reduce the torsional vibration) and improve lubrication/oil flow in the synchronizer components. Churning loss is a speed dependent loss that will produce higher gearbox drag at higher speeds as large quantity of oil has to be moved per time unit [11]. The oil resists movement of parts, there by dampening vibrations and mechanical shocks. The oil volume in the gearbox was 5 l. The volume of the oil was increased to 7 l to provide higher splash in to wear gap, better lubrication and to dampen the vibrations. The 5 l CFD analysis showed less oil movement around the 3rd and the 4th synchronizer rings. The 7 l CFD analysis revealed that more oil reached the friction lining material on the 4th gear side via the openings of clutch body ring lugs. The opening of clutch body ring lugs in the 3rd gear side was covered by the larger 3rd gear, so less quantity of oil was reaching the lining material as shown in Fig. 26, Fig. 27. The 4th counter shaft gear had larger diameter and had higher immersion depth on the oil compared with the 3rd gear. It drew up more oil and created more splash and stirred the oil bath more [13]. In the TV 440
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Q (L/min)
B. K, et al.
0.9 0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1 0 0
0.02
0.04
0.06
0.08
0.1
0.12
0.14
0.16
h (mm) z=4
z=5
z=6
z=7
Fig. 25. Volume of Oil required for full wetting of gear cone surface for different gear cone to synchronizer ring gaps
Fig. 26. Oil flow distribution on the 3rd & the 4th gear positions for 5 l and 7 l of oil (section taken between synchronizer ring and CB ring).
test for 65 h, the 3rd synchronizer had a wear of 0.52 mm in 5 l gearbox and the wear reduced to 0.3 mm in 7 l gearbox. For the 4th gear the wear was 0.5 mm in the 3rd gear and the wear reduced to 0.2 mm in the 7 l oil gearbox as shown in Fig. 28. This wear reduction happened due to the higher oil level and higher oil churning. The increased level of oil volume increased the drag of counter shaft gears and increased the drag torque on synchronizers [14]. This reduced the free movement of the synchronizer rings by dampening the vibration transferring from the 5th gear to counter shaft gears and in turn reduced the impact in all gear synchronizer positions. 4.2.6. Engine dyno test The improvements carried out in the study such as clutch diameter increase, reduced inner synchronizer ring clearance, oil grade and oil volume were validated in the engine dyno test. The initial design had a wear of 0.9 mm in the 3rd synchronizer position on the engine dyno. This wear reduced to 0.1 mm with 441
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Fig. 27. Oil flow distribution on the 3rd & the 4th gear positions
0.6
Wear (mm)
0.5 0.4 0.3 0.2 0.1 0 1
2
3
4
Gear Clearance 0.5mm, SAE80W90, 5L Oil
Clearance 0.5mm, SAE80W90, 7L Oil
Fig. 28. Influence of oil volume on synchronizer wear
the 10% increased clutch diameter, 0.5 mm inner synchronizer ring clearance, with 7 l of SAE80W90 oil as shown in Fig. 29. The wear was reduced in all gear synchronizer positions. 5. Conclusion This paper describes the synchronizer ring wear due to torsional vibration in manual transmission gearbox synchronizers. The experimental study reduces the synchronizer wear. The conclusions from the study and the experimental results are as follows:
1
Wear (mm)
0.8 0.6 0.4 0.2 0 1
2
3
4 5 6 Gear Std. Dia Clutch, Clearance 3mm, SAE80, 5 Ltr oil 10% increaesd Dia Clutch, Clearance 0.5mm, SAE80W90, 7 Ltr oil Fig. 29. Configuration of engine dyno test and synchronizer wear 442
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1. The torsional vibration causes non-uniform wear on the synchronizer ring. The inner ring wear is higher than the outer ring wear. 2. The transfer of torsional vibration into the gearbox is dampened by the bigger diameter clutch. The inner synchronizer ring guide restricts the free movement of the ring. Further high viscosity oil dampens the vibration which cause impact of synchronizer ring and wear. 3. More amount of oil is getting churned with the increased oil volume, which pushes more oil in to the synchronizer ring and dampen the vibration. The increase in the drag on the counter shaft gears further cut the vibration transfer path. 6. Future Scope of work Increase of clutch size increases the downshift reflected inertia. Oil volume and the oil viscosity increases the drag of the gearbox. These will affect the gear shift quality, which can be solved by optimizing the synchronizer index percentage, tooth thickness and tooth chamfer angles. Nomenclature A b dmax Fs Grpm Hrpm h rm IRef Nc Q u(z) Z α ∆rpm ω ωa ωi
Surface area (mm2) width of synchronizer ring (mm) Max cone diameter (mm) Force at sleeve (N) Gear speed (rpm) Hub speed (rpm) radial space between the gear cone and synchronizer ring (mm) Mean radius of synchronizer ring (mm) Reflected inertia of the gear (Kgm2) Number of cones Axial volumetric flow rate (mm3) Axial velocity (mm/s2) Full oil film length (mm) Cone Angle (°) Difference in speed (rpm) Angular acceleration (rad/s2) Angular acceleration of the synchronizer ring / hub (rad/s2) Angular acceleration of the gear (rad/s2)
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