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Improvement in Gear Shift Comfort by Reduction in Double Bump Force of Passenger Vehicles.

Introduction

Manual transmissions have come a long way since their inception. We had constant mesh gearboxes where the drivers had to be skillful enough to shift smoothly. Now, we have synchronized transmissions where one does not need to be skillful to achieve a smooth shift but as the competition is so intense in current automotive sector, even synchronized transmissions have to be amongst the best if we want to remain competitive. There has been immense focus on customer touch points, namely Steering, Gearshift lever, Clutch and overall ergonomics which have come to defining the overall perception of a vehicle. Gear shifting is a continuous event and any small discomfort causes customer dissatisfaction. Hence, manufacturers are always trying to achieve that elusive perfect gearshift feel to gain competitive advantage.

Gearshift quality analysis (GSQA) is an objective way of measuring the parameters involved in gear shifting process. GSQA relates the subjective perception of customer about gear shifting comfort into the technical objective parameters. Static GSQA measures different parameters like selection travel, shift travel, end stop force and free play. Shift system contributes to most of static GSQA parameters. Dynamic GSQA measures different parameters like synchronization (Sync) force, double bump (DB) force, synchronization integral, and double bump integral. Gearbox layout and synchronizer geometry contributes to most of the dynamic GSQA parameters, which are simple to predict through calculations.

There are different phases involved in gear shifting; one of the important phase is 'Synchronization&'. The synchronizers are complex in nature because it is having dynamic interaction with gear shifter, drive train and the clutch. Double bump is the second peak force after synchronization force. Higher double bump is perceived as an irritant during shifting. It is most difficult to predict or control as it is a dynamic event and depends on the response of the entire system, including the selector mechanism, transmission, driveline, and synchronizer system. Effect of double bump is measured in GSQA by two important parameters namely DB to Sync force ratio and number of peaks in DB region.

There are few papers on optimization of double bump and these are - 'Use of detent profile in reducing double bump [1]&', Asymmetrical chamfer on gear engagement ring and shifter sleeve teeth [2]&', 'Use of stiffness optimization of shift forks [3]&' and 'Use of inertial mass to mask double bump [4]&'. However, there is no comprehensive literature available on double bump reduction. The objective of the paper is to focus on double bump, its effects and solutions. The paper describes various design parameters affecting the double bump such as synchronization capacity, strut gap, wear gap, spline clearance, linkage ratio and gearbox drag etc. Detailed analysis of components contributing towards gearbox drag was carried out and it was observed that use of lapping on synchronizers helps to reduce drag.

The paper also outlines the simulation model built to study effects of various design parameters as well as system parameters. Simulation model was used to study effect of driveline system such as axle shaft and propeller shaft stiffness on double bump. In few cases such as north-south gearbox layout and mechanical shift linkage where double bump is inevitable, masking double bump is the only solution. The paper suggests masking techniques which result in damping of double bump resulting in better shift feel, though this results in compromise in other areas of shift feel.

Synchronizer Basics

There are several detailed literatures on basic synchronizers. Socin Walter [5], Basics of synchronizers by Hoerbiger [6], have explained the synchronizing process in detail. There are six different phases [8] of synchronization are shown graphically in fig.1 in terms of shift force versus shift time.

1. Disengagement Phase

During this stage, the gear knob is moved from one gear position to neutral gear position. This can be seen as a positive force peak denoted by A&' as shown in fig.1, which is nothing but the in-gear detent force, which has to overcome for disengaging gears. Additionally, it also has a component which depends on the back taper angle and torque on the particular gear. Back taper creates a swallow force, which also has to be overcome during disengagement. The acceptable values are between 15 to 20N for disengagement.

2. Neutral Phase

The detent now moves into the neutral detent groove hence reducing the detent force. The shifter sleeve chamfers are now completely out of the previous gear position.

3. Pre-Synchronization Phase

The sleeve chamfer and synchro cone teeth have to contact each other in order to generate synchronization force else, the sleeve might pass through without synchronizing resulting in clash engagement. The carrier ensures this orientation. As soon as the shifter sleeve is axially displaced, synchro cone is pressed by carrier thereby achieving two things; one is to wipe the oil on the cone surface to increase the friction coefficient of cone, secondly to orient the synchro cone to index position.

4. Synchronization Phase

The synchro cone is now in blocked or indexed position and further movement of the sleeve increases the friction torque on synchro cone until the gear and sleeve speeds match each other. The maximum peak force observed between start of synchronization phase (2) and end of synchronization phase (3) during synchronization process is the synchronization force. Maximum synchronization force is as indicated in fig.1 as 'C&'. It depends on the synchronizer cone diameter, material combination with gear and cone angle and synchro/sleeve chamfer angles.

5. Free Flight Phase

The speed of sleeve and speed gear are now synchronized. This causes the cone torque drop to zero. At this stage, the axial force on sleeve is more than the cone torque hence the sleeve rotates the synchro cone and the sleeve can now pass through, to engage with gear engagement teeth. During this phase, the gear might again start to slow down due to effect of drag.

6. Final Engagement Phase

The shifter sleeve chamfers now contact the gear engagement teeth, giving rise to a second peak called double bump or second bump. Maximum double bump force is as indicated in fig.1 as 'D&'. Further travel of shifter sleeve results in end stop, which indicates gear shifting is complete. At this position, the shifter sleeve chamfers get lock with back taper of gear engagement ring.

Double Bump Phenomenon

There are two peak forces during stage (2) and (5) within a short time interval refer to fig.1. The second peak force occurring after synchronization force is called double bump, in fact the name double or second bump comes from here. After synchronization the synchro cone torque reaches zero and at this moment the shifter sleeve axial force i.e. index torque is greater than cone torque, hence the cone deindexes. The shifter sleeve now passes through to engage with engagement ring chamfers to turn the engagement ring to complete the shift. This force when shifter sleeve teeth hit the engagement ring chamfers is called double bump. Hence double bump is an integral part of the synchronization process. Typical Double bump force on shift force vs shift time graph is shown below in fig.2.

As the instantaneous position of gear engagement ring teeth chamfer and shifter sleeve chamfer is indeterminate, hence analyzing double bump is a bit complex issue. Sometimes, shifter sleeve teeth do not experience any impact resistance by gear engagement ring teeth due the probability of shifter sleeve teeth coming exactly opposite to tooth space between gear engagement ring. At this condition double bump peak is clearly missing which is indicated in below fig.3.

Double bump force is analyzed in GSQA by two different parameters as described below:-

1. DB to Sync Ratio

The ratio of double bump (DB) force to synchronization (sync) force commonly referred as 'DB / Sync ratio&'. In passenger cars, double bump peak force should be half of peak synchronization force or lower i.e. this ratio should be less than 0.5 The driver is likely to perceive double bump only if its more than half of synchro force or more than 20N .

2. No of Peaks in Double Bump Region

The number of peaks in Double bump region is also an important factor. The number of double bump peaks represent number of attempts made by shifter sleeve to engage with gear engagement teeth on a speed gear. When shifter sleeve passes through the free flight phase, it interacts with gear engagement ring chamfers to produce a second peak force. Sometimes shifter sleeve bounces many times before engaging into the gear engagement teeth chamfers in double bump region. In this case speeds of shifter sleeve and engagement ring are synchronized but then the relative speed picks up after free flight stage due to various reasons. If peak is more than one, then it is termed as Nibble. Nibble may consist of two or three peaks with varying handball force intensity as shown in below fig.4. Nibble is perceived by the driver as kick back feel at the knob along with grating sound.

Forces Experienced in Double Bump

Double bump force has two force components, one is impact force (Fi) and other is sliding force (Fs).

Double bump force(F) = Impact force(Fi) + Sliding force (Fs)

Impact Force (Fi)

Amount of double bump force depends on shifter sleeve teeth chamfer, gear engagement ring chamfer, and their relative positions as shown in figure.5. After synchronization when shifter sleeve approaches towards gear engagement ring, the first contact made by sleeve chamfers with engagement ring chamfers which produces impact force (Fi).

Mathematically,

[mathematical expression not reproducible] Eq. (1)

Where

Fi is Impact force in N

I is Synchronized inertia in Kgm2

[DELTA]n is speed difference between shifter sleeve and gear engagement ring.

[alpha] is sleeve teeth chamfer angles in degrees.

R is Pitch circle diameter of shifter sleeve in m.

T3 is time in sec

M is mass in kg

Vo is shift speed in m/sec.

[eta] is shift system efficiency

Lr is Leverage ratio or linkage ratio.

Sliding Force (Fs)

After first engagement of shifter sleeve teeth chamfer with gear engagement teeth chamfer, the shifter sleeve teeth must slide over gear engagement teeth chamfer to reach the final engagement position. The reaction force due to this sliding between chamfers creates sliding force (Fs) and it depend on teeth chamfer angle, coefficient of friction as shown in fig.6.

Mathematically,

[mathematical expression not reproducible] Eq. (2)

Where

Fs is Sliding force in N.

[empty set] is wrap angle in degrees.

mfr is friction of synchronizer in Nm.

T4 is time in sec.

[micro]c is coefficient of friction between shifter sleeve and gear engagement teeth chamfers.

[micro]s is coefficient of friction between coupling sleeve and synchro hub.

Parameters Affecting Double Bump

Double bump is not only affected parameters inside gearbox but also by system level parameters. Detailed fishbone diagram is attached in 'Appendix I&'. The fishbone diagram identifies many possible causes for an effect or problem. The important parameters identified and thoroughly analyzed are discussed below:

Design Parameters

1. Synchronizer capacity Synchronizer capacity plays an important role in double bump. Due to higher synchronization capacity, rate of synchronization is very steep which results into large torque wind up downstream of the synchronizer. As the synchrocone deindexed during the free flight phase, it accelerates the gear engagement ring. Reducing the synchronization capacity will reduce the rate of synchronization; intern reduce the torque wind up which minimizes double bump.

2. Strut Gap Strut gap is the clearance between the strut and synchronizer outer cone face when shifter sleeve is in neutral position. If strut gap is too large then the detent load is not that effective leading to notchy feel. In addition, if strut gap is too small then the synchro cone is continuously wiping oil from gear cone, which increases the drag. This increased drag increases speed difference between gear engagement ring and shifter sleeve and increases double bump force. Therefore, it is necessary to have optimum strut gap in design in order to have reduced double bump force.

3. Synchronizer Wear Gap Wear gap is the clearance between the gear engagement ring and synchro cone. If the wear gap is too high, then distance travelled by shifter sleeve during free flight is increased. Longer free flight results in higher speed difference due to drag and leads to issue of double bump.

4. Spline Clearance The spline radial clearances between synchronizer hub and shifter sleeve is critical to ensure smooth sleeve travel. Higher clearance leads to stick-slip phenomenon and lower clearance result in jamming the shifter sleeve. Hence optimum hub to sleeve teeth clearance results in lower double bump.

5. Gearshift Lever Ratio Increasing linkage ratio helps in reducing the impacting force component of double bump. Since higher impact force will cause the shifter sleeve to bounce on gear engagement teeth chamfers instead of turning the gear engagement ring. However, the linkage ratio also needs to be optimization based on aspect ratio i.e. ratio of shift travel to select travel of shift mechanism and the force felt by driver.

6. Detent Force It comprises neutral detent force, strut detent force and shifter-rail detent force. Shift rail detent has both positive and negative force during shifting process. The negative force is responsible for the swallow force experienced while shifting. The swallow force helps to reduce the double bump. The negative force should start exactly at the free flight to best utilize the detent force to reduce double bump.

7. Shifter Fork

Material. The shifter fork plays a very important role in double bump reduction. The idea is to store energy in the shifter fork through deflection during synchronization phase and release the same during free flight phase much like a spring being compressed during synchronization phase and released during free flight. This is the reason a lot of OEM&'s were using brass forged shifter forks as shown in fig 7. Hence, Automobile manufacturers are now switching to fine blanked sheet metal forks as shown in fig 8.

Stiffness. Jaideep Singh&'s paper [3] on shifter fork correlation to gear shift quality describes in detail the effect of stiffness on double bump. Simulation results are also available which display a reduced double bump force perceived at the gearshift knob.

The stiffness of the shift fork has major role in improving shift feel as it directly relates to double bump and synchronizer nibble. During synchronization, when blocker ring is indexed, an extremely stiff fork limits the amount of energy stored in shift fork, a lower stiffness shift fork will reduce the free flight time resulting in less speed difference between gear engagement ring and shifter sleeve which happens due to drag. Influence of the fork stiffness on the double bump and nibble was studied and trials were conducted with reduced /optimized stiffness values from 1000N/mm to 5,000N/mm and then the results were compared as shown in fig.9

The results show reduction in DB/Sync force ratio with lower fork stiffness.

Contact Area. Fork contact with sleeve also plays also plays key role in reducing double bump such as using a full contact on shifter fork on sleeve as shown in fig.10. It is not recommended as the fork deforms and in turn tilts the sleeve while engaging. Whereas, shifter fork with 3 pad as shown in fig.11 has two-point support with a third support to restrict amount of sleeve tilt. It is a good design where the sleeve has a limited allowable tilt, which can adjust the sleeve on impact with gear engagement teeth.

Detent Profile Optimization. Rohit Kunal&'s paper [1] on Simulation of Gear shift force Curve and shift Rail Ramp Profile explains in detail the procedure and formula to optimize detent profile for best utilization of detent force. Basic idea is to use detent swallow force in overcoming double bump force such that the driver doesn&'t feel the double bump force.

Synchronizer Geometry. Asymmetrical chamfers can be used to address this issue as described by Amit Sandooja in his paper on "Automotive Synchronizer with Asymmetric Tooting" [2]. The main advantage of Asymmetrical chamfer is to reduce the free flight zone hence reducing effect of drag on gear to be synchronized.

Transmission Drag

As seen from the equation for impact force refer eq. (1), Impact force increases with increase in speed difference after synchronization i.e. [DELTA]N. A higher impact force causes sleeve to bounce back rather than turn the gear for final engagement, hence leading to double bump phenomenon. If the [DELTA]N is higher, it can also lead to multiple bumps resulting in a phenomenon called Nibble as explained before. The rpm difference arises after synchronization due to the friction drag of the gearbox, which slows down input shaft post synchronization.

The transmission drag depends on a number of components as described below:-

* Bearings

* Gears in mesh

* Synchronizer

* Oil used in transmission

* Oil seal friction

* Shim selection

The drag on gearbox has two components, resistance due to oil splash and friction. Oil has two parameters impacting drag. One is the churning effect that depends on the oil quantity and the other is the oil viscosity. The other part of drag depends on friction, which comes from bearing pre load like in taper roller bearings, friction from synchro cones on speed gears, friction from oil seals and gear mesh friction.

The drag has an adverse effect during down shifting and actually assists gear shifting during upshifting. The drag can be reduced using a low viscous oil such as synthetic oils, optimizing strut load on synchronizers, opting for ball bearings instead of taper roller or cylindrical bearings and optimizing oil splash or churning.

Experiments were carried out on a transmission to know exact contribution of components/parameter on drag torque. All measurements were carried out at neutral condition. Drag measurement results as indicated in table.1 and table.2 with different condition as shown below: -

It was observed that contribution of synchronizer is more towards drag of the transmission than the oil seal and bearing. Contribution of various components on drag is shown below in pie chart refer fig.12

To reduce drag, synchronizer lapping process was added, its effect on drag torque was measured, and the results are as shown in table. 3

The results show that lapping process helps to reduce drag torque by 15-20% without any change to other transmission components and which intern helps in reducing double bump.
FIGURE 12 Contribution of transmission of components towards drag torque

Gear + Shaft + Bearing  G+S+B         17%
Oil seal                Oil seal       8%
Synchronizer            Synchronizer  58%
Oil                     Oil           17%

Note: Table made from pie chart.


System Interaction - Drive Train

Drive train in this context refers to all parts right from gearbox input shaft to vehicle tires; including propeller shaft, differential, axle shaft, flange etc. More components connected to each other reduces the stiffness of the system irrespective of the type of joints used.

When we compare a front wheel drive fig.13 layout with a rear wheel fig.14 layout.

It can be seen that front wheel layout consists of fewer parts as compared to rear wheel drive layout. The output shaft or drive shafts are directly connected from Transaxle&'s Gear final drive splines to the wheel hub. Whereas the rear wheel drive setup consists of a propeller shaft which is connected through splines to gearbox output shaft, a slip joint on propeller shaft to accommodate the wheelbase change during bump and rebound condition, differential unit and finally axle shafts connecting to the wheel hub.

As it can be seen the rear wheel drive has a lower stiffness compared to front wheel drive. Hence during the drive transmission, the drivetrain has to deform rotationally as per the stiffness before transmitting torque. During gear shifting when the clutch is disengaged the drivetrain tries to unwind since the torque from engine is absent. The unwinding of drivetrain will result in a continuous torsional vibration before stabilizing; this will cause an angular to and fro motion which can cause the sleeve to bounce back and results into double bump.

Hence, double bump can be minimized by optimizing the Drive train stiffness in Rear wheel drive vehicles and obviously double bumps are lower or less probable on front wheel drive vehicles.

Masking Double Bump

Where Double bump is inevitable we can reduce the perceived double bump using masking techniques like adding compliance in the system. Typically, North -South transmission layouts which have direct shift mechanism or a remote shift linkage system have severe double bump perception since linkage transfers even slightest change in force or vibration inside gearbox to the driver

1. Inertia Mass on Shift Lever Inertia mass can be used on shift levers to lower double bump force. This also works on the Fork stiffness principle, difference being here, the inertia stores the force thereby taking care of double bump. The increase in weight also gives a definite shift feel to the driver however addition of too much weight makes the feel sluggish and hard since the weight has to be moved across the shift stroke.

As per the above equation when the shift mass increases, the impact force on sleeve increases. Now, since the impact force is coming from the inertia mass, the driver doesn&'t perceive the double bump on gear knob. Manish Kumar Sharma & Jinesh Savla&'s paper [4] on Shift system inertia mass optimization techniques to minimize double bump for manual transmission gives detailed information on the procedure for optimizing inertia mass. Basically, this is a design of experiments(DOE) with addition of shift mass to check the subjective gear shift feel as well as tracking double bump values to achieve a tradeoff between good shift feel and reduced double bump values. Addition of weights are seen to be easy for transaxles rather than in inline transmissions. For Transaxles, weights can be directly added on shift levers as shown in fig.15 Inline transmission requires an additional linkage for adding mass as shown in fig 16.

2. Use of Heavy Knob Using a heavy knob also helps in reducing double bump perception but on the flip side excess weight on the knob might also results in sluggish shift feel hence a tradeoff between shift feel and double bump is to be established. It is also important to note that increasing knob weight in case of mechanical linkage is sensitive for gear jump out hence adding weight has to be carefully evaluated in testing and also while tracking the shift feel improvement.

3. Rubber Damper on Shift Lever General approach is to design a telescopic shaft with rubber in between to damp the double bump force as shown in fig 17. The compliance will help in reducing double bump perception but deteriorate the shift precision. Hence a tradeoff of shift precision and double bump perception is to be established. Adding compliance also deteriorates the selection feel making it rubbery and sluggish which will not only give a bad selection feel but will also result in difficulty finding desired gates during selection.

The above problem is solved by adding the rubber differentially with thicker section along shift direction and thinner section along selection direction. Hence retaining good selection precision and also dampening double bump during gear shifting. Fig.18 shows the differential compliance in shift and selection direction on gear shift lever.

4. Optimizing Stiffness - For Rod Linkage On vehicles with remote linkages an additional mounting is used to anchor the remote linkage on cab or gearbox through an 'I Bolt&' as shown on fig 19. Optimizing the stiffness of damper between I bolt and its housing also contributes to reduction in double bump perception.

5. Optimizing Stiffness - For Rod Linkage In case of Transaxles with cable shifts the cable end fittings have dampers on both sides as well on the end terminal which connects to the gearbox levers. All these dampers as shown in fig.20 and fig.21 can be tuned by adjusting the rubber thickness or shore hardness to achieve the best results.

Double Bump - Model Simulation

Predicting gear shifting mechanism behavior on a particular vehicle is the need of today&'s advance system. Simulation model was derived in accordance with Ricardo.

In this dynamic model, all components that have degree of freedom are modelled as if they are point mass (translational degree of freedom) or point inertia (rotational degree of freedom). Flexible components modelled to have stiffness and a damping coefficient but no mass or inertia. In cases where a component's mass, inertia, or stiffness is an influence on system behavior, then the component modelled as a combination of these two types. The most obvious example of this is in the case of shafts, such as drive shafts and propeller shafts. In this dynamic model, positive rotation of transmission and driveline components defined as the direction in which the components nominally rotate during a normal gearshift. All rotating components therefore have a positive rotation at the start of simulation. Torques acting on such components defined as being one of the following: -

* 'Drive Torque&' - Acts to accelerate component

* 'Load Torque&' - Acts to decelerate component

The basic simulation model divided into 5 sections: -

1. The primary input to the model is the motion of the gearshift lever in 'External Selector&'. The external selector is formed by the actuator, the gearshift lever and the external shift shaft.

2. The motion further transfers to 'Internal selector&'. The internal selector is consisting models of rails, fork and shifter sleeve. The internal selector system acts on the synchronizer models, which is located within 'Base Transmission&'.

3. The rotating parts of the transmissions are modelled within 'Base Transmission&'. It contains model of input and output shafts, gears on these shafts and synchronized gears freely rotating on these shafts.

4. This 'Base Transmission&' system interacts with 'Clutch&'. As the clutch is disengaged during the gearshift, dynamics of the outer assembly (including friction plates) and torque in torsional spring is considered while modelling.

5. This transmission system interacts with 'Driveline&'. The driveline model includes the propeller shaft, final drive ratio, differential, and the stiffness, inertia and backlash in the driveshaft&'s and CV joints. The model is given angular velocity of the wheels as an input calculated from the engine rpm and the initial gear. This velocity is maintained as constant, as if the vehicle&'s speed maintained.

Below is basic simulation model architecture shown in fig.22

Model provides output in terms of actual values related to GSQA parameter of -

* Synchronization force in N.

* Double bump force in N.

* Number of peaks in DB region.

* DB to sync force ratio.

* Synchronization integral in Nm.

* DB integral in Nm.

* Total shift time in sec.

In addition, model provides output as simulation, where we can run through real time synchronization process for selected or particular gear shifting. Simulation of shifter sleeve travel through synchronizer cone and gear engagement ring as shown fig.23 along with it plots the handball force (shift force) against shift time for selected shift. We can plot this graph against sleeve travel too.

This kind of simulation helps to predict the behavior of synchronization well in advance than the having traditional approach of carrying GSQA on a vehicle level and verify the synchronizer parameters. This also helps to carry out and verify effect of design parameters affecting synchronizer performance and intern gearshift quality. It also predicts effect of system level parameter like differential mechanism, propeller shaft and axle shaft on gearbox and its quality.

Results

We instrumented one vehicle with GSQA measurement unit and data captured for upshifts like 1-2, 2-3, 3-4, 4-5, 5-6 and downshifts like 6-5, 5-4, 4-3, 3-2, 2-1. While capturing data vehicle run for 30 times in each individual shift in order to have consideration to all relative positions of shifter sleeve w.r.t. synchrocone and gear engagement ring. Result parameter related to double bump i.e. number of peaks in double bump region and DB/Sync ratio.

Result parameter related to double bump are number of peaks in DB area and DB/Sync ratio. For a particular shift, high values and low values of these parameter plotted and mean value denoted by point. All such mean points joined indicating mean values for these parameter for all upshifts and downshifts, which indicated in Fig.24 and Fig.25. From this graphs it observed that double bump is more prominent and higher in upshifts than the downshifts. In addition, upshifts has wider range for DB/ sync force ratio for upshifts than the downshifts.

Basic simulation model output results for the same gearbox for gear shift from first gear to second gear is as shown in fig.26. The results from simulation were compared to actual GSQA results and it showed a very good co-relation, thereby validating the simulated model.

It seen from above upshift chart, double bump force is more than the synchronization force and observed DB/ Sync ratio is greater than 0.5, which is not good for a shift quality. It predicts the high and low values for double bump as observed on vehicle. Similar results were run for different upshift like 2-3, 3-4, 4-5, 5-6 and downshift like 6-5, 5-4, 4-3, 3-2, 2-1 shifts which indicated in fig.27 and fig.28 under 'Base design&' line joining points. It shows that dynamic models deemed accurate against GSQA measurement performed on actual vehicle.

The following changes were done to base design to reduce double bump

* Number of synchro cone reduced.

* Lower module spline teeth used for shifter sleeve, synchrocone and gear engagement ring.

* Synchrocone lapping introduced to reduce drag.

* Shifter shaft detent profile modified to match initiation of force drop at free flight zone.

* Carrier and sleeve profile modified

* Wear and strut gap optimized.

* Linkage ratio increased and free play reduced for rod shifting arrangement.

Simulation model was run with the above changes and results showed improvement in double bump which are indicated in Fig.30 and Fig.31

Improvements from base design changes shows improvement towards double bump to synchronization force and number of peaks in double bump region. Help us to reduce double bump.

Our experience in this field says that sometimes parameters other than gearbox level also leads to have impact on gearshift feel. This is due to system interaction of synchronizer with other subsystems. Model help us to verify the effect of various such subsystems on gearshift feel.

Increasing Axle Shaft Stiffness

Simulation was run with increased axle shaft stiffness and it was found to have significant effect on double bump. When axle shaft stiffness was increased by 30% results showed 50% reduction on double bump force. Physical trials were conducted on vehicle and perceivable reduction in double bump was felt inline with what was predicted in simulation model.

For gear shifting from first gear to second gear, model results with existing / unchanged axle shaft stiffness is shown in Fig.29 and Fig.30 shows the simulation results showing reduction in Double bump force after changing axle shaft stiffness by 30%

Increasing Propeller Shaft Stiffness

Another important subsystem of driveline, which has effect on gearshift feel, is propeller shaft. Like axle stiffness, simulation trials were done to increase the propeller shaft stiffness by 30% and verify its effect along with increased axle shaft stiffness. For gear shifting from first gear to second gear, model results with combined increase in axle shaft and propeller shaft stiffness is shown in Fig.31. Below results from dynamic model indicates that double bump peak was further minimized with stiffer propeller shaft.

To summarize, the effect of stiffness increment over modified design on DB/Sync ratio is shown in Fig.32. Combining both, changes i.e. increment in axle shaft stiffness by 30% and propeller shaft stiffness by 30% indicates double bump/ sync force ratio reduced by 65-70 %

Conclusion

There has been a phenomenal increase in the expectation on gear shift quality and double bump has been a constant pain area especially for North- South gearboxes with remote or direct linkages. There is no comprehensive literature on double bump which the paper intends to accomplish by enlisting various parameters affecting double bump. Outline of program architecture is also described for virtual simulation of gear shifting event. Results of simulation showed excellent correlation with actual GSQA results.

Detailed study was done to understand contribution of various components on drag. Synchronizers were found to have a significant contribution for which Synchronizer Lapping was introduced hence reducing drag by 2Kgcm. Though the effect of increasing powertrain stiffness is explained which reduces double bump by 60-70% the same needs to be further studied for concequential effect on critical speeds, NVH and resonance with natural frequencies of other parts on vehicle.Lastly masking techniques have been discussed for situations where Double bump is inevitable for example in North-South transmissions with direct or remote shift linkage. The masking strategy is useful to reduce the intensity of Double bump though it may be noted that masking deteriorates the shift feel and a tradeoff is required between crisp shift feel and double bump intensity.

Contact Information

Santosh Deshmane, Tata Motors Ltd

santosh.deshmane@tatamotors.com

Onkar Gangvekar, Tata Motors Ltd.

onkar.gangvekar@tatamotors.com

Samson Rajakumar, Tata Motors Ltd.

r.samson@tatamotors.com

Definitions/Abbreviations

Fig - Figure

DB - Double Bump

GSQA - Gear Shift Quality Analysis

Sync - Synchronizer

DOE - Design of experiment

References

[1.] Kunal, R., Adiga, G., Gill, S., and Sharma, M., "Simulation of Gear Shift Force Curve and Shift Rail Ramp Profile," SAE Technical Paper 2010-01-0896, 2010, doi:10.4271/2010-01-0896.

[2.] Sandooja, A. and Kunal, R., "Automotive Synchronizer with Asymetric Toothing," SAE Technical Paper 2011-01-0724, 2011, doi:10.4271/2011-01-0724.

[3.] Singh, J., Verma, A., Kunal, R., Balpande, A. et al., "Shifter Fork Stiffness Correlation to Gear Shift Quality," SAE Int. J. Commer. Veh. 6(2):498-509, 2013, doi: 10.4271/2013-01-2447.

[4.] Sharma, M. and Savla, J., "Shift System Inertia Mass Optimization Techniques to Minimize Double Bump for Manual Transmission," SAE Technical Paper 2012-01-1999, 2012, doi: 10.4271/2012-01-1999.

[5.] Socin, R. and Walters, L., "Manual Transmission Synchronizers," SAE Technical Paper 680008, 1968, doi:10.4271/680008.

[6.] Hoerbiger, " Basics of Synchronizers", http://mobile.hoerbiger.com/upload/file/2013 basicsofsynchronizers.pdf, Accessed June, 2017.

[7.] Hackl, Thomas, Werth,Peter, Hiraiwa, Kazuyoshi, "Applicability of The New Kyowa High Performance Synchronizer", Technical Paper, F2006P097T, Hofer Powertrain & Kyowa Metal Works CO., LTD, https://www.hofer.de/de/download/F2006P097T-Paper.pdf

[8.] Gangvekar, O. and Deshmane, S., "Grit Blasting on Synchronizer - To Resolve Early Crashing Complaint," SAE Technical Paper 2017-01-1769, 2017.

Appendix I. Fishbone diagram for Double Bump Phenomenon

Santosh Deshmane, Tata Motors Ltd.

Onkar Gangvekar and Samson Rajakumar, Tata Motors Ltd.

History

Received: 12 Jan 2018

Published: 08 Oct 2017

e-Available: 08 Oct 2017

Citation

Deshmane, S., Gangvekar, O., and Rajakumar, S., "Improvement in Gear Shift Comfort by Reduction in Double Bump Force of Passenger Vehicles," SAE Int. J. Passeng. Cars - Mech. Syst. 10(1):2018, doi:10.4271/06-11-01-0006.

doi:10.4271/06-11-01-0006
TABLE 1 Effect of components on drag torque.

                                   Drag torque at neutral
Sr.No  Measurement condition         condition in Kgcm

1      Fully built Transmission              12
2      All Synchro cone removed               5
       from transmission
3      All oil seals removed from             4
       transmission

TABLE 2 Effect of oil quantity on drag torque.

                                   Drag torque at neutral
Sr.No  Measurement condition         condition in Kgcm

1      Transmission with full oil            12
2      Transmission without oil               9

TABLE 3 Effect of synchronizer lapping process on drag torque.

                                   Drag torque at neutral
Sr. No  Measurement condition        condition in Kgcm

1       Fully built Transmission             12
        with regular synchronizer
2       Fully built Transmission             10
        with lapped synchronizer
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Author:Deshmane, Santosh; Gangvekar, Onkar; Rajakumar, Samson
Publication:SAE International Journal of Passenger Cars - Mechanical Systems
Article Type:Report
Date:Jan 1, 2018
Words:5942
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