POLITECNICO DI MILANO

Transcription

1 POLITECNICO DI MILANO School Of Industrial and Information Engineering POLO REGIONALE DI LECCO Master s Thesis Suspension Design Feasibility Study of Light Commercial Vehicle in ADAMS/Car Supervisor: Prof. Francesco Braghin Company supervisor : Mr. Vincenzo Abbatantuoni Master Thesis by: Sandip Kumar Mat. No October

2 DECLARATION I hereby declare that this thesis, submitted to Politecnico di Milano as partial fulfillment of the requirements for the degree of Master is completely novel and has never been presented at any other University for an equivalent degree. I also certify that the document below has been exclusively done by me, with the exception of certain standardized data and technique, the sources f o r which are appropriately cited in the references. This thesis may be made available within the university library and may be photocopied or loaned to other libraries for the purpose of consultation. October 2015 Sandip Kumar 2

3 ABSTRACT In the present era of advancement in science and technology, computer aided engineering plays a pivotal role in automotive industry. Now complete vehicle can be modeled and simulated for various road conditions. Results have been obtained with high degree of accuracy for the same. There had been very good correlation with the real time test data and simulated results. On some occasions simulation results had a better accuracy while compared to the real test data. This boosted the confidence of industry and now, more and more tests are performed with the help of CAE. Following the same trend in this project a complete multibody model for a small passenger taxi car was developed in Adams/car prior to building the real prototype. The modeling was done in accordance with the dimensions obtained from CAD model. This ensured model was well with in design parameters. After completion of the model, it was first evaluated at subsystem level extensively to see the design conformity and eliminate any observed variation. Parallel and opposite wheel travel analysis was performed to simulate bump, rebound and roll condition. After that full vehicle simulations were performed to assess straight line stability, lane change performance and cornering behavior. All the results and outputs were discussed in detail with the possibility of future application and works. 3

4 ACKNOWLEDGEMENT I take this opportunity to express my sincerest gratitude to my university supervisor Prof. Francesco Braghin for his expert guidance and freedom of work he provided me throughout the project. His constructive criticism and valuable suggestions motivated me to perform better than my capabilities. I would also like to thank my company supervisor Mr. Vincenzo Abbatantuoni for giving me this wonderful opportunity. His extensive support, patience and trust in my work kept me going in difficult times and helped me in successful completion of the project. Along with this my gratitude goes to Mr. Luca Marano for his expert comments and suggestions. His knowledge and understanding of vehicle dynamics always inspired me to learn more. A special thanks to Mr. Testi, Mr. Catelani and Mr. Brutti for their support and guidance with Adams software. I would also like to extend my gratitude to Mr. Marchisio, Davide and Antonio for helping me with performance parameters and design files. I would like to thank my parents, jaji, my sisters Sarika, Nitu and my dear friend Ms. Rubina Mahtab for their constant encouragement and support throughout this arduous journey, without which this would not have been possible. Last but not least a special mention to my extraordinary friends Dhanush, Shehzad, Arijit, Bhuvan, Mukund, Shyam, Pradip, Ashish, Maddy, Geo, Vikas, Shahnawaaz, Amrita and Gesu who never failed to boost my morale and encouraged me throughout this journey. Thank you all. 4

5 LIST OF FIGURES Figure 1 Double wishbone suspension system Figure 2 McPherson strut Figure 3 Semi-trailing arm suspension Figure 4 Ackerman Geometry Figure 5 Camber angle Figure 6 Toe angle Figure 7 Caster angle Figure 8 Kingpin and scrub radius Figure 9 Vehicle coordinate system in accordance with SAE convention Figure 10 Rear suspension assembly Figure 11 Final assembly top view Figure 12 Final assembly front view Figure 13 Final assembly front ISO view Figure 14 Integration of MSC. ADAMS with CAE software Figure 15 Steps of suspension analysis Figure 16 MacPherson modelling component topology Figure 17 Front suspension links Figure 18 Steering links Figure 19 Front suspension assembly with test rig Figure 20 Rear suspension assembly with test rig Figure 21 Lumped mass body system Figure 22 Engine and powertrain system representation Figure 23 Front and rear tyres Figure 24 Database for full vehicle assembly Figure 25 Full vehicle assembly front view Figure 26 Full vehicle assembly side view Figure 27 Full vehicle assembly top view Figure 28 Full vehicle assembly Front-Iso Figure 29 Compliance matrix Figure 30 Camber change Design of Experiment Figure 31 Driving machine function Figure 32 Vehicle straight line test set-up Figure 33 Single and double lane change Figure 34 Add text Figure 35 Ramp steer test parameters Figure 36 Event builder

6 LIST OF TABLES Table 1 Specifications and target dimensions Table 2 Suspension parameters Table 3 Type of joints and their respective degrees of freedom Table 4 Front parts and joints topology Table 5 Static parameter variable table Table 6 Steering system joints topology Table 7 Rear assembly joints topology

7 LIST OF GRAPHS Graph 1 Power train map Graph 2 Observed toe with sign change Graph 3 Toe adjustment geometry Graph 4 Modified toe with negligible sign change observed Graph 5 Caster and KPI v/s wheel travel curve Graph 6 Front track change v/s wheel travel Graph 7 Rear track change v/s wheel travel Graph 8 Steer angle vs rack displacement Graph 9 Turn radius Graph 10 Ackerman error Graph 11 Longitudinal vs lateral displacement Graph 12 Front steering angle Graph 13 Inclination angle vs time Graph 14 Lateral force vs time Graph 15 Chassis displacement Graph 16 Yaw rate and steering wheel angle at 50kmph Graph 17 Yaw rate and steering wheel angle at 55kmph Graph 18 Yaw rate and steering wheel angle at 60kmph Graph 19 Steering wheel angle v/s lateral acceleration Graph 20 Turning radius v/s steer wheel angle Graph 21 Yaw rate v/s steer wheel angle Graph 22 Velocity v/s steer wheel angle Graph 23 Understeer curve

8 TABLE OF CONTENT DECLARATION 2 ABSTRACT 3 ACKNOWLEDGEMENT 4 LIST OF FIGURES 5 LIST OF TABLES 6 LIST OF GRAPHS 7 1. INTRODUCTION LITERATURE REVIEW Introduction and Role of suspension system in a vehicle Types of suspension system Double wishbone suspension system: McPherson strut suspension: Semi-trailing arm rear axle: Suspension system parameters Ackerman Camber Toe Caster Kpi and scrub radius DESIGN OF SUSPENSION COMPONENTS AND GEOMETRY Choice of type of suspension system Front suspension and steering system Rear suspension system Approach to CAD modeling and assembly of components Final representation of CAD model MODELING AND ANALYSIS IN ADAMS Introduction to ADAMS/Car Working principle Approach to system modeling Global reference coordinate Local coordinate system Markers Approach to system modeling 26 8

9 4.4.1 Standard user interface Template builder mode System modeling Modeling of front suspension assembly Modeling rear suspension assembly Body and chassis system Powertrain assembly Front and rear tyres modelling Full vehicle assembly Evaluation of suspension and steering characteristics Definition of complex matrix Wheel travel analysis Steering analysis Full vehicle simulation Straight line maintain test Single Iso-lane change maneuver Ramp steer Conclusion Appendix Front assembly hardpoints Steering system Rear suspension system Body subsystem Powertrain subsystem Tire property file. 62 REFERENCES 68 9

10 1. INTRODUCTION The company wanted to launch a new short wheel base passenger taxi car capable of seating 4 persons including driver with a maximum payload of 450 Kgs. The vehicle was expected to be compact in size with the sufficient ground clearance, affordable and durable enough for rough driving conditions. To make it compact track width, wheel base and ground clearance was fixed according to the market survey and benchmarked value. So, the main goal was to develop a multibody model of the planned vehicle and analyze the suspension behavior prior to build real prototype. This would show a broad picture of the real scenario and road conditions. Any required modification or change in geometry, components would be possible to do at design stage itself making project efficient in terms of time and money. 10

11 2. LITERATURE REVIEW 2.1 Introduction and Role of suspension system in a vehicle The role of a car suspension is to maximize the friction between tires and the road surface to provide steering stability with good handling and to ensure the comfort of the passenger. If the roads were perfectly flat then there is no need of suspension system but unfortunately this is not the case. The role of suspension varies between commercial and race cars. While the race car suspension system demands high performance without any compromise to squeeze out minimum possible lap time, commercial vehicles require maximum life time and durability of components with compromise in performance. The main aim of suspension system in commercial vehicle is to have lowest amount of tire wear, minimum noise and vibration with maximum ride comfort. A good suspension system should provide best possible ride and handling performance, which is only possible if wheel follows the road profile with very little tire fluctuation. The vehicle must be in steerable condition at all times and driver should get a good response while maneuvering. This factor is ensured by the fact that vehicle responds favorably to the forces generated by the tires during cornering or accelerating with a good dive and roll geometry design 2.2 Types of suspension system In commercial or any other vehicle, chassis is the mainframe and must be able to handle the varying engine power, acceleration, peak cornering speeds at all times which leads to the choice of independent suspension system. Well known governing factors for this choice is low weight, no mutual wheel influence, little space requirement, easier steerability, a kinematic or elasto-kinematic toe-in change and ease of adjustment. Low weight and no mutual influence on the wheel are two important characteristics for good road handling on uneven road surface with curves. Considering the above factors, independent suspension system was chosen. While we speak of all the possible design and performance requirements there comes the engineer s nightmare to achieve all this within given constraints which is daunting. At various stages of design, compromise had to be made to achieve a close to ideal performance. The design constraints have been outlined by the problem statement for suspension system of small, lightweight, affordable and appealing passenger car. The choice of suspension system and its components are guided mainly by required investment, manufacturing cost, packaging constraints and performance benchmark values. All the design constraints and bench marked values will be discussed and explained throughout the project Double wishbone suspension system: Figure 1 Double wishbone suspension system 11

12 It consists of two transverse links (control arms) on either side of the vehicle which are mounted to rotate on the frame and connected on the outside to steering knuckle via ball joints. The greater the effective distance between the transverse links, the smaller the forces in the suspension control arms and their mounting becomes, i.e., the deformation in the component is smaller and wheel control is more precise. Double wishbone suspension has great advantage with the kinematic possibilities. The inclination of control arm can decide the height of body, roll and pitch while varying the length of the same influence the angle movement of the compressing and rebounding wheels i.e. the camber and track width change. With all the advantages in terms of vehicle performance and ease of adjustment this type of suspension system is not suited for the vehicle under study. The main reason is requirement of larger control arms with high inclination which is not possible with such a small track width and large cabin space requirement. It would be extremely difficult to achieve a proper geometry with shorter arms and packaging space McPherson strut suspension: Figure 2 McPherson strut This type of suspension system is further development of double wishbone suspension. The upper control arm is replaced by a pivot point on the wheel panel, which takes the end of the piston rod and the coil spring. Forces from all the directions are concentrated at this point. The main advantage of the McPherson strut is that all the suspension components can be combined into one assembly. The steering knuckle can be welded, brazed or bolted firmly to the outer tube. Further advantages are lower forces in the chassis side of the mounting, long spring travel and larger packaging space, as there is no upper control arm. Despite minor disadvantages like lack of noise reduction and bulky structure, this type of suspension system fits right in choice because of previously mentioned advantages and company s experience in designing and manufacturing this kind of suspension. Hence McPherson was chosen for front suspension of the vehicle. 12

13 2.2.3 Semi-trailing arm rear axle: Figure 3 Semi-trailing arm suspension The semi-trailing arm rear suspension is characterized by balanced comfort and driving behavior. With the semi-trailing arm rear suspension, the wheels are mounted on links that move at an angle to the vehicle s longitudinal axis as they deflect and rebound. The compact design also allows for a large luggage compartment. When wheel goes in bump and rebound-travel they cause spatial movement, so the drive shafts need two joints per side with angular mobility and length compensation. The horizontal and vertical angles determine the roll steer properties. Camber and toe-in changes increase, the bigger the angles are. Semitrailing axles have an elasto-kinematic tendency to oversteer. With the ease of driveshaft mounting and available wheel travel it was selected for rear suspension system. 2.3 Suspension system parameters The kinematics of the suspension system can be visualized as a body moving in space relative to another body with three components of translation and three components of rotation. While the independent type suspension system allows for relative motion between the wheel and the vehicle body without affecting the other wheel, it has only one path of motion. Like any other single body, wheel has six degrees of freedom in space out of which five dof s are restrained by the suspension linkages. These linkages also severely limit the orientation of the wheel as it travels in jounce and rebound against spring and damper which can rotate about its three axes due to the geometry of the suspension. These rotation deviate the geometry from the ideal suspension design. At maximum, the designer can try to offset the effect of such rotations with fine tuning the geometry but still it would result in change in camber, caster and toe-angle. An insight to the same has been briefly discussed below Ackerman When front wheel drive vehicle is steered away from straight-ahead position, the design of steering linkage determines whether the wheels stay parallel or one wheel steers more than the other. This difference in steer wheel angle is the effect of Ackerman geometry and the device which provides this effect is called Ackerman steering. There is no four bar linkage mechanism which can give a perfect Ackerman geometry, 13

14 but a close to perfect condition could be achieved. The condition to meet this criterion is that vehicle must be driving slowly and free from any lateral forces. Since the passenger cars are designed for low lateral acceleration use, it is preferred to use Ackerman geometry. The Ackerman condition is expressed by the following formula: Where: is the steer angle of outer wheel is the steer angle of inner wheel w is the trackwidth l is the wheelbase Figure 4 Ackerman Geometry Camber Camber angle is the angle that the wheel plane makes with the vertical axis. It is positive when the top of the wheel leans outwards, away from the vehicle body and negative when it is inwards. Cornering force of the tyre depends on its angle relative to the road surface. Hence camber is a major contributing factor for good road grip. When the tyre moves on the road the rubber is elastically deformed, as the tread is pulled through the tyre/road interface, which causes an additional lateral force known as camber thrust. Due to the contribution of this camber thrust, tyre develops its maximum lateral force at a small camber angle. Hence it is suggested to provide a small amount of camber angle in the direction of wheel rotation to optimize tyre performance during the turn. 14

15 Figure 5 Camber angle Passenger cars are designed with soft roll stiffness to provide a smooth ride. This low roll stiffness results in large wheel travel causing large camber change and eventually reduced tyre performance and excessive wear. Therefore, within the given parameters a compromise must be made to achieve a balance between the two, which has been discussed during suspension analysis Toe It is the angel between the longitudinal axis of the vehicle and the line of intersection of the wheel plane and the vehicle XY plane. It is positive if the wheel front is rotated towards the vehicle body and vice versa. Figure 6 Toe angle It can be expressed in degrees or radians but it is more common to express it as the difference between the track width measured at leading and trailing edges of the tyre. Toe control is important as it directly affects three major performances, i.e. corner entry handling, straight line stability and tyre wear. When a car is running in a straight line, the wheels in the given axle should directly point ahead, for maximum power and low tyre wear. Directional stability of the vehicle is increased by toe-in, while toe-out increases the steering response. Too much of toe-in causes rapid wear at the outer edges of the tyre and vice versa. 15

16 2.3.4 Caster Caster angle is the angle in the side elevation (vehicle XZ plane) between the steering (kingpin) axis and the vehicle axis. It is positive when the top of the steer axis is inclined rearwards and vice versa. Figure 7 Caster angle A negative caster helps in quick steer return, increases the straight-line ability of the vehicle, provides a large tyre contact patch area during turn and good steering feel. But if the caster is increased to a large value it will also cause an undesired increase in steering effort Kpi and scrub radius The Kingpin inclination (KPI) or steering axis inclination (SAI), is the angle formed between vertical and the line joining the upper and lower ball joints (steering axis). It affects the self retuning mechanism of the steering to straight ahead position, bringing the wheel to its highest point. Positive kingpin is defined with the upper ball joint closer to the chassis than the lower. Kingpin is also used to set scrub radius, which can be adjusted to zero value when the steering axis coincides with the tyre centerline on the ground. But excessive KPI and high steering angles lead to positive camber change and corner weight variations. Figure 8 Kingpin and scrub radius Scrub radius, also known as kingpin offset, is the lateral distance between the centerline of the wheel and intersection of the kingpin axis at the ground. A large scrub radius means reduced steering effort at the cost 16

17 of increased friction which causes the tyre to scrape through the surface while turning, as the tyre no more turns along the center line. So it is desirable to minimize the scrub radius in order to make the steering less sensitive to road irregularities, braking etc. A zero scrub radius eliminates the negative effects on the steering but will also leave the driver with a dead feel of the steering. of Acceptable scrub radius is thought to be those smaller than 25% of the tread width. 3. DESIGN OF SUSPENSION COMPONENTS AND GEOMETRY Before proceeding to the design of suspension components and related geometry, a comprehensive market survey was done by the company to find out the customer requirements and market needs. Then based on the inputs from the survey benchmarking process was carried out. The basic requirement of the vehicle was it should be low cost, light weight, spacious and full fill the needs of south east Asian countries. Proceeding with this idea, first size and general specification of the vehicle was decided as follows. Length Width Height Ground clearance Kerb weight Wheel radius Engine Transmission 3 m 1.5 m 2.5 m m 450 KG 12 inches 230cc water cooled Manual, Rear wheel drive Table 1 Specifications and target dimensions. Next a decision had to be made for the track width and wheel base of the vehicle. For this one constraint was already defined as length and width of the vehicle. Second was influenced by the fact that interior of the vehicle must be spacious along with the provision of luggage compartment. So a wheel base of 1.8m and track width of 1.2 m was proposed. Keeping in mind the road irregularities a target ground clearance of m was also decided. These three values were subject to minor changes as the design would progress. 3.1 Choice of type of suspension system During this process a comparative study on various types of suspension system for front and rear was done. Reasoning for the same has been discussed in the following section. 17

18 3.1.1 Front suspension and steering system As the wheel base was fixed and it had been stated that interior should spacious then it became quite obvious that front driver and passenger seats would be located close to the wheel envelop. Considering these facts, there were two good possible choices. It was either double wishbone suspension or McPherson strut type suspension system. Double wishbone suspension required two transverse links (control arms) on either side of the vehicle which would be mounted to rotate on the frame and connected on the outside to steering knuckle via ball joints. This kind of suspension provided good control over geometry and improved behavior in dynamic condition. But major disadvantage came in terms of packaging issue. Placement of upper arm required a large wheel envelop and that would cramp the driver and front passenger leg space. It also required more components (upper A-arm, extra bushings and ball joints for mountings) and assembling time. In terms of ergonomics and cost it was not a perfect choice. This ruled out any scope of using double wishbone type suspension. Second best choice for front was McPherson strut type suspension system. This is a further development of double wishbone suspension as top. A-arm is replaced by the strut system. It only required lower control arm mounting and placement of strut in vehicle body full filing the criteria of large cabin space. It also needed less number of components and assembling time, so in terms of cost also it was an advantage to use this. The final factor which dominated this choice was company s previous experience in this type of suspension and readily available assembly components and no need for any new investment. Hence for front, McPherson strut suspension was decided. For steering system rack and pinion type steering system was selected keeping in mind its simple and effective design. This type of steering also provided a good possibility of having good Ackerman geometry. Assembly and mounting would also take less space Rear suspension system During the market survey and benchmarking process choice of engine was also locked. Vehicle had to be rear wheel drive to meet the norms. Rear suspension needed to be spacious enough to accommodate engine and provide some space for luggage compartment too. Passenger ride comfort was also a major driving factor. This required a suspension design with softer springs. After a comparative study of solid axle, quadlink, semi trailing arm type of suspension it was decide to go with semi trailing arm kind suspension as it provided a balance between comfort and driving behavior. It also met the criteria of low cost and assembling time. Solid axle kind would have required mounting of a separate differential and it was neither suitable for engine nor for the light weight nature of vehicle. 3.2 Approach to CAD modelling and assembly of components. Once finished with the type of suspension it was time to prepare the CAD model to evaluate and tune the suspension as per the requirement. With the company s vast database and experience in CAD modelling, assembly process was quite convenient. Most of the component design and drawings were available in the database. Whole vehicle design and assembly was carried out according to SAE convention. According to SAE convention all points of interests were described as coordinates dimensioned from the intersection of the zero planes in the three-dimensional reference system. XYZ coordinates were dimensioned to their respective planes. 18

19 Figure 9 Vehicle coordinate system in accordance with SAE convention After fixing vehicle coordinate system assembly process started with fixing the wheel center and then moving on to hub, knuckle and strut assembly. Static parameters like camber and toe were set to zero with the option to change at later stage. It was decided to go with stock steering knuckle so the only choice of altering or controlling kingpin inclination, scrub radius and caster was dependent on strut top mount. Hence according to the requirement following suspension parameters were decided. SUSPENSION Tire Static Loaded Radius m Camber Angle 0.0 deg Kingpin Inclination Angle 16.0 deg Scrub Radius m Caster Angle 4.0 deg Caster Trail Rack Stroke STEERING m m Inner Steer Angle at Lock 50.4 deg Outer Steer Angle at Lock 37.7 deg Max. Ackerman Error 3.75 deg Rack Force Ratio at Lock Table 2 Suspension parameters 19

20 The 12" wheel had an offset of 40 mm (ET) which allowed a better placement for the tie rod outer ball joint inside the rim. The brake drum was positioned accordingly. In order to use the identical knuckle geometry received from the database, the outer ball joint center location of the lower control arm was also modified. Figure 9. Front suspension assembly Taking into account, the tire SLR of 261 mm and the ground clearance target of 240 mm, the lowest design position for the LCA inner pivot axis was kept at -95 mm in Z-axis initially. The ideal position of the inner lower control arm attachment points was fixed on a line on the Y-Z plane, which passed through the virtual swing arm rotation center of the MacPherson-type strut suspension. This ensured that the suspension bump-steer was minimized. This was rechecked during toe control analysis. For rear suspension the trailing arm pivot radius was kept 354 mm in order to accommodate the tire R12 tire. In order to achieve sufficient amount of rear axle understeer during steady-state and yaw damping during transient maneuvers, the pivot axis inclination was tuned to provide toe-in with increasing suspension compression as per the preliminary idea. Engine mounting points were decided based on the fact that CG should stay low and sufficient ground clearance is also available. Drive shaft inner joint was also one of the factor while deciding engine location as care was taken to keep CV joint angle to minimum for efficient power transfer. In static condition CV joint had an angle of deg. The tentative mounting of spring and dampers were according to the packaging space which was subject to change as per the suspension analysis. 20

21 Figure 10 Rear suspension assembly 3.3 Final representation of CAD model. Figure 11 Final assembly top view 21

22 Figure 12 Final assembly front view Figure 13 Final assembly front ISO view 22

23 After completing the full vehicle assembly in CAD, hardpoints were exported for modelling and analysis of suspension design in Adams/car. 4. MODELING AND ANALYSIS IN ADAMS 4.1 Introduction to ADAMS/Car Adams, developed by MSC Software Corporation is a multibody dynamics software that is widely used in engineering industry. Adams/car is the part of Adams software suite and it provides a specialized environment for modeling vehicles. The software formulates equations of motion based on absolute coordinates to obtain a time response of the system. It can be time consuming as complex assemblies often involve large systems of nonlinear differential algebraic equations requiring large amount of computing power. Adams/car allows the user to create and test virtual prototypes of the vehicle subsystems and complete vehicles much like the physical system. An extensive library of macros is also built into the program to speed up the model creation. Using Adams, full vehicle assembly can be created /modified rapidly and can be simulated for various conditions to understand their performance and behavior. Based on the analysis result suspension geometry, spring rates and other kinematics can be altered in no time to get the desired performance behavior. Since the main goal of this project is to design and study the suspension system and vehicle behavior before building prototype, Adams was selected to perform this task. Prior to modelling suspension system in Adams, an extensive study in vehicle dynamics, vector theory and classical approach of designing was performed to understand the working of the software. It is of utmost importance to understand what goes behind the screen and the working principle of this software. During the study and modeling of the system it was found that if the user knows the working of the software then it becomes very easy and efficient to modify the system and obtain desired results. 4.2 Working principle The main analysis code consists of a number of integrated programs that perform three dimension kinematic, static, quasi-static or dynamic analyses of the mechanical system. These programs form the core of the solver. In addition to these a number of other programs are linked to the core solver which is used to model vehicle tire characteristics, automatically generates vehicle suspension geometry etc. Once the model is defined the core solver assembles the equation of motion and solves them. Possibility of inclusion of differential equation in the solution makes it easier to model various control system. Complete Adams working process has been explained in diagram. 23

24 Figure 14 Integration of MSC. ADAMS with CAE software The first step for simulation is to prepare a data set which defines the system. It includes rigid parts, connecting joints, motion generators, forces and compliances. Adams provides the user to use various joints and connectors used in the system. For the real time similarity it is necessary that each rigid body s mass, center of mass location and moment of inertia is well defined. For simpler bodies Adams itself calculate all the above mentioned requirements. For some complex components it was calculated in CAD and then manually entered in the software. Further each body has a co-ordinates system which can be defined in local co-ordinate system or global co-ordinate system. Parts and bodies move according to this definition during simulation. The relative motions between different parts in the system are constrained using joints, gears, couplers etc. The next step is defining the external and internal force elements. External forces can be constant, time dependent function or any other state dependent function. These forces can be translational or rotational. Internal forces act between two parts like spring, damper or rubber mounts. These are referred as action and reaction forces and they always produce equal and opposite forces on two parts connected by the force element. Adams allows user to effectively access any displacement, velocity, acceleration or other force when defining the force equation. Forces can also be switched on or off during simulation progress. Precaution must be taken to ensure formulations are continuous in time domain to avoid any problem during the 24

25 numerical solution of the resulting equations. If these forces or any other parameters are not in the given range the solver is not able to execute the equation of motion resulting in failed simulation. In Adams, vehicle suspension bushings and joints are represented by a set of six action and reaction forces, which hold the two parts together. The equations of force are linear and uncoupled. Following the common notion too much complexity is bad thing, Stiffness of the bushings and joints were mostly left unaltered. A list of joints with their degrees of freedom has been updated in the following table. Constraint element(joint) Translational constraints Rotational constraints Coupled constraints Total constraints Cylindrical fixed Planar Rack and pinion Revolute Spherical Translational Universal Coupler Table 3 Type of joints and their respective degrees of freedom As per the real conditions, in Adams also, forces through the road are transferred to the tire. For each tire on the model, Adams calculate the three orthogonal forces and torques acting at the wheel center as a result of the condition at the tire road surface contact patch. It is resolved at the wheel center and then software integrates it through time to find the new position and orientation of the vehicle and repeats the process. In short suspension analysis can be explained with following diagram. Figure 15 Steps of suspension analysis In Adams suspension analysis is divided in two steps i.e., pre-processor and post processor. During first step all input data is given through GUI or by command files and then solver performs the analysis. Post processing is the second step in which all the results files are viewed. Post processor software is the part of Adams suite and is used with other applications too. Post processor made it easy to understand the vehicle 25

26 behavior. While looking to the model in motion it was easy to debug any anomaly. Post processor also provided the option to perform mathematical operations and statistical analyses on plot curves. All the results, simulations and presentation of curves were done in post processor. 4.3 Approach to system modeling Since, it is quite clear that Adams/car is a powerful tool for simulation not modelling so, before starting the modelling in Adams, it is highly suggested to sketch out a system schematics which would typically illustrate the items such as parts, joints, imparted motion applied forces and its location. It also helps in predicting degrees of freedom and develop a basic understanding how system will work. For the same reason detailed CAD drawing was prepared and it made the modelling work less tedious to model the system in Adams. Next step is to set up the reference frame. For a multibody three dimension description is required. It not only set up the configuration and physical properties of the model but also to describe the calculated outputs such as the displacement, velocities and acceleration. For the given model three types of reference frame coordinate system was used Global reference coordinate This is the single inertial frame that is fixed and at rest. Any point defined to this reference frame has zero velocity and acceleration. The ground reference is fixed on a body. For a single suspension model the ground part may be taken to encompass the points on the vehicle body or sub-frame to which the suspension linkages are attached. During a full vehicle modelling the ground part was related to the surface of road to formulate the contact forces and moments in the tire model. It can be also considered as the origin of the entire model. All other reference frames are measured relative to the ground reference frame Local coordinate system Each body or part has a local co-ordinate system which moves or changes according to the orientation of the part. Adams takes it as body co-ordinate system. It is defined relative to the ground reference frame Markers These are the points located in the model to define the entities like mass center, position, spring ends etc. It can belong to part or ground. Orientation of marker is important. For example in case of revolute joints, marker should be along the axis of rotation or else joint will not behave as intended. The Euler angle method can be used for orientation of the same. Basic modeling components in ADAMS Basic components are rigid bodies (part), geometry (marker), constraints (joints, gears etc.), forces (applied, spring forces etc.), user defined algebraic and differential equations. Part statement will be used to define rigid body or lumped mass. Suspension components like control arms, wheel, steering knuckle etc. will be modelled as rigid bodies. For dynamic analysis full information like center of mass, moment of inertia, orientation etc. is required. 4.4 Approach to system modeling Adams works in two interfaces namely: 26

27 4.4.1 Standard user interface In this mode standard templates are available (most commonly used designs) with all the required configuration, constraints and joints. It also facilitates the option of modifying various components, hard points and several other parameters to create the desired geometry Template builder mode In this mode user need to enter all necessary data for the part to be created like role type, hard points, joints, constraints, mass and inertia properties. It takes more time and effort than standard interface to create desired template. It can be avoided if the type of geometry to be used is less complicated and more common. But for creating a new geometry with desired properties and parameters it is always preferred. For this particular study of the suspension system, choice was predefined as MacPherson for front and Semi trailing arm for rear. Reasons for the same have been discussed in design of suspension section. Both Macpherson and semi trailing arm suspension templates are available in Adams/car database. So it was decided to go with standard user interface and modify the hard points, mass-inertia properties etc. to obtain the desired configuration and geometry System modeling Steps to create a system model in Adams: Creation of database and then selecting the working directory. Then select sub assembly or assembly from file-new-menu To obtain the complete vehicle model following assemblies and sub-assemblies were modeled and then finally assembled: Front suspension assembly Rear suspension assembly Body and chassis Powertrain assembly Front and rear tires Modeling of front suspension assembly Front suspension assembly consisted of 3 subassemblies: Suspension links with springs and damper, steering system and test rig Modeling MacPherson suspension link subassembly for front The template represented a standard design of one piece lower control arm and subframe. Upright was represented by a combination of links to which hub, lower control arm, tie rod and strut was mounted. Foreaft and lateral motions of the uprights were regulated by the lower control arm. Steering rotation of the upright was controlled by the tie rod and vertical, side, front view rotations were controlled by the strut. During quasi-static analysis a static rotation control actuator locked the degree of freedom of the hub. All the above mentioned parts were connected with different types of joints and bushings which has the following topology explained in the table. Main part Connecting part Type of joint. Lower control arm Subframe Revolute joint Upright Lower control arm Ball joint Upright Upper strut Cylindrical joint 27

28 Tie rod Upright Spherical joint Tie rod tie rod to steering (mount part) Convel joint Subframe Subframe to body (mount part) Fixed joint Upper strut Strut to body (mount part) Hook joint Spindle Upright Revolute joint Table 4 Front parts and joints topology Figure 16 MacPherson modelling component topology After defining the topology, parameter variables were defined to incorporate any camber or toe in static condition. In this case it was set to zero. Parameter variable value unit Camber 0 Deg. Toe 0 Deg Table 5 Static parameter variable table For creating the assembly following steps were followed: Step 1: Creating subassembly. a) File New Subsystem b) subsystem name : (Front_link_geometry) 28

29 c) Minor role : Front d) Template name : (Acar database _MacPherson.tpl file.) Subassembly appeared on the screen. Step 2: Modifying subassembly Subassembly was modified using the hardpoints obtained from the CAD geometry. Refer Appendix 6.1 a) Select Adjust Hardpoints Table. b) Following hardpoints were entered in the table. (As given in appendix) c) Spring preload and damper properties were also adjusted as per the requirement. The following geometry was obtained: Figure 17 Front suspension links Modeling steering system subassembly The type of steering to be modelled was rack and pinion. In the model rotatory motion of the steering wheel was translated to linear motion by the pinion gear. The rack moved tie rods back and forth to steer the vehicle. The motion from the steering to pinion was transferred by steering shafts which was connected with a series of hook joints. Lower column shaft to the rack housing was connected by a revolute joint. Shaft to the pinion was connected by a torsion bar bushing. Pinion to the rack housing was connected with the revolute joint. During kinematic mode, a reduction gear was active which connected the steering input shaft revolute joint to the pinion revolute joint. The motion of rack to rack housing was constrained by a translation joint. In case steering assist was required, it was given by VFORCE. Steer assist VFORCE was controlled by steer_assis_.tbl file of the Adams/car database. In case steer assist was not needed it was switched off. All parts and connection and joints topology can be summarized in the following table. 29

30 Main Part Connecting part Type of joint Steering column Intermediate shaft Hook joint Intermediate shaft Steering shaft Hook joint Rack Rack housing Translational joint Steering wheel Steering column to body (mount part) Revolute joint Pinion Rack housing Revolute joint Steering column Steering column to body(mount part) Cylindrical joint Steering shaft Rack housing Revolute joint Rack housing Rack housing mount(switch part) Fixed joint Table 6 Steering system joints topology During analysis, switching between kinematic and complaint mode was carried out by the parameter variable. It was set by the hidden option under parameter variable steering_assist_force. Maximum values of steering angle, rack displacement, rack force and steering wheel torque was set using the same option. Steps for creating steering subsystem has been described in following steps. Step 1: Creating subassembly a) File New Subsystem b) Enter subsystem name : (Steering_system) c) Minor role: Front d) Template name: (Browse to Acar database and the select steering.tpl file.) Subassembly appears on the screen. Step 2: Modifying subassembly Subassembly was modified using the hardpoints obtained from the CAD geometry. a) Select Adjust Hardpoints Table. b) Following hardpoints were entered in the table. (As given in appendix 6.2) c) Steering gear ratio was also modified from the default value. The following geometry was obtained. 30

31 Figure 18 Steering links Creating front suspension assembly Once both front suspension links and steering subsystem was created and saved in template files, front assembly was created in following steps. a) File New suspension assembly. b) Enter Assembly name (front_suspension_complete) c) Suspension subsystem (Browse and select Front_link_geometry file from subsystem.) d) Select Steering subsystem (Browse and select Steering_system file from subsystem.) e) Suspension Test Rig (select _MDI_SUSPENSION_TESTRIG) and apply. 31

32 Figure 19 Front suspension assembly with test rig Modeling rear suspension assembly Rear suspension assembly consisted of 2 subassemblies: Suspension links with springs and damper, driveshaft system and test rig. The model to be created was semi trailing arm type suspension. It was non steerable Modeling rear suspension link subassembly In this model left and right trailing arms were connected to the rigid subframe and in turn subframe was connected to body mount part through bushings. The wheel center was located by the arms. Spring and damper act between arms and body mount parts. The rotational degree of freedom of the hub was locked by the static rotational control actuator during quasi-static analysis. Joints and connection topology has been summarized in the following table. Main part Connecting part Type of joint Strut to body (mount part) Upper strut Hook joint Arm Subframe Revolute joint Subframe Subframe to body (mount part) Fixed joint Lower strut Arm Hook joint Upper strut Lower strut Cylindrical joint Spindle Arm Revolute joint 32

33 Tripod Tripod to differential(mount part) Translational joint Spindle Arm Revolute joint Table 7 Rear assembly joints topology For rear assembly too parameter variables were decided as explained in previous section for front suspension. Static camber and toe values were kept zero. Complete model was created in following steps. Step 1: Creating subassembly a) File New Subsystem b) Enter subsystem name: (Rear_link_geometry) c) Minor role : Rear d) Template name: (Browse to Acar database and the select trailingarm.tpl file. ) Subassembly appears on the screen. Step 2: Modifying subassembly Subassembly was modified using the hardpoints obtained from the CAD geometry as follows. a) Select Adjust Hardpoints Table. b) Following hardpoints were entered in the table. (As given in appendix) Creating rear suspension assembly Once the rear link template file was created and saved in database, complete assembly with the test rig was created in following steps. a) File New suspension assembly. b) Enter Assembly name (Rear_suspension_complete) c) Suspension subsystem (Browse and select Rear_link_geometry file from subsystem.) Suspension Test Rig (select _MDI_SUSPENSION_TESTRIG) and apply. 33

34 Figure 20 Rear suspension assembly with test rig Body and chassis system Body and chassis system was modeled as single lump of mass which consisted of structural mass of body-inwhite, engine, exhaust system, fuel tank, vehicle interior, driver, passenger and any other payload. Weight distribution was done by editing trim part mass. The lump mass chassis had the limitation in terms of torsional stiffness as it could not be defined in this kind of model. It required making several rigid bodies and then connecting them with spring and damper to obtain the desired stiffness. Since evaluation chassis stiffness was not the main goal of this study, it was left unchanged. Chassis template was created in following steps. Step 1: Creating subsystem. a) File New Subsystem b) Enter subsystem name: (Chassis_body) c) Minor role: Any d) Template name: (Browse to Acar database and the select rigid_chassis_lt.tpl file.) Subsystem appears on the screen. Step 2: Modifying subassembly Subsystem was modified using the hardpoints obtained from the CAD geometry as follows. a) Select Adjust Hardpoints Table. b) Following hardpoints were entered in the table. (As given in appendix) 34

35 Figure 21 Lumped mass body system Powertrain assembly Powertrain assembly was modeled using standard template available in the database after modifying the mounting points. Connection of engine to the drive shaft was verified and changed to chassis from ground. In graphical representation it only shows the mounting point and drive shaft connection. Engine properties were left unchanged except changing the mass. Graph 1 Power train map 35

36 Power train template had a very simple topology because it is just a functional representation of the engine. The only rigid parts along with the engine body were differential outputs and revolute joints which connected the rigid bodies to the engine body. Powertrain template was created in following steps. Step 1: Creating subsystem a) File New Subsystem b) Enter subsystem name: (Power_train) c) Minor role: Rear d) Template name: (Browse to Acar database and the select Powertrain_lt.tpl file. ) Subsystem appears on the screen. Step 2: Modifying subassembly Subsystem was modified using the hardpoints obtained from the CAD geometry as follows. a) Select Adjust Hardpoints Table. b) Following hardpoints were entered in the table. (As given in appendix) Figure 22 Engine and powertrain system representation Front and rear tyres modelling Three basic functions were being provided by the tire system. It supported vertical load, developed lateral forces for cornering and longitudinal forces for acceleration and braking. The template consisted of wheel parts rigidly connected to mount parts. The tire contact patch forces were transformed in forces and torques applied at the hub. Force calculations were done by user defined sub routine files based on the tire properties. Tire topology was defined by a fixed joint which connected the wheel part to the spindle mount part. For better and ease of handling Pacejka2002.tir tire model was used. It was based on special version of magic formula. It described the tire behavior for a smooth ride of frequency up to 8 Hz. It was also suitable for a speed up to 30 Mph. Tire property file used has been added in the appendix. 36

37 Figure 23 Front and rear tyres Full vehicle assembly Once all the subsystems were created and saved in the respective database, full vehicle assembly was done. Steps: Creating full vehicle assembly a) File New Full vehicle assembly. Required fields were filled as per the database templates shown below and the applied. Figure 24 Database for full vehicle assembly 37

38 Figure 25 Full vehicle assembly front view Figure 26 Full vehicle assembly side view 38

39 Figure 27 Full vehicle assembly top view Figure 28 Full vehicle assembly Front-Iso 39

40 4.5 Evaluation of suspension and steering characteristics In Adams/car during suspension analyses, a total of 38 characteristics (Camber, toe, roll, caster etc.) are calculated. Force limit for the left and right test rig jack is -2.0e+04 and 4.0e+04 N. This force limit can be modified while working in the template builder mode, using define actuator option. The suspension and steering characteristics that Adams/car computes are based on the suspension geometry, suspension compliance matrix or both. Suspension geometry refers to the position and orientation of suspension parts relative to ground as the suspension is articulated through its ride, roll and steer motion. Suspension compliance matrix refers to incremental movements of the suspension due to the application of incremental forces at the wheel center. Throughout the motion, at each position Adams compute the compliance matrix. Characteristics such as suspension ride and camber aligning torque are the result of compliance matrix Definition of complex matrix Compliance matrix is the partial derivative of displacement with respect to applied forces. If the system is assumed to be linear, then its movement can be prescribed with the applied force. Here, matrix element Cij is the displacement of system, degree of freedom I due to a unit force at degree of freedom j. Adams uses a 12x12 matrix relating the motion of the left and right wheel centers to unit forces and torques applied to the wheel centers. It has the following form: Figure 29 Compliance matrix Further for calculating characteristics such as camber, caster, scrub, caster moment arm etc. Adams uses steering axis of the suspension. User has two method available i.e geometric method and instant axis method. 40

41 Geometric method: The steer axis is calculated by passing a line through the selected points. Generally it is suitable for solid axle suspension. Instant axis method: To calculate the steer axis at a given position, Adams/car first locked the spring travel and applied an incremental steering torque or force in all directions. Then from the resulting translation and rotation of the wheel carrier part instant axis of rotation for each wheel was calculated. For this particular suspension instant axis method of calculation was selected Wheel travel analysis The wheel travel analysis allowed to see, how the suspension characteristics changed throughout the vertical range of motion. Adams/car provided option of three kinds of wheel analysis i.e parallel, opposite, single wheel analysis. First two were sufficient to evaluate the characteristics and has been discussed in detail in following sections Parallel and opposite wheel travel analysis These wheel travel analysis were performed to check the behavior of vehicle kinematic parameters when both wheel went in jounce and rebound condition simultaneously or in opposite direction. Opposite wheel motion simulated body in roll condition. Main factors which were under investigation were change in camber, toe, kingpin inclination, and track radius and track width. Before running test, the parameters like tire model, tire stiffness, wheelbase, sprung mass, CG height, wheel mass were defined. Once the suspension parameters were fixed, analysis parameters were defined. During the analysis, the test rig applied forces or displacement or both to the assembly as defined in a load case file. Adams/car generated a temporary load case file based on the input given which was used for future simulations. Test parameters: Suspension assembly : Final_front / Rear_assembly Output Prefix : PWT_front / PWT_Rear Number of steps : 150 Mode of simulation : Interactive Vertical setup Mode : Wheel center Bump Travel : 80 Rebound travel : -60 Travel Relative To : Wheel center Control mode : Absolute The camber change observed for front in the beginning with the base geometry was in access of ±3.2º. This high camber change was not acceptable as that would lead to large chassis roll and loss of tire contact patch area, leading to excessive tire wear. It needed modification, so a design of experiment was carried out using Adams/Insight to find out the most critical factors. After the experiment, it was observed that Z-component of lower front control arm pivot point and lower control arm outer ball joints were the major contributing factors, evident from the following figure. 41

42 Figure 30 Camber change Design of Experiment Finally a camber change of 2.4º was achieved. The change was still large so a compromise was done to maintain the target ground clearance, because changing the Z-component of lower control arms would move the vehicle further close to the ground. Further a change in sign for toe was also observed which would result in unfavorable side drift/jerk during jounce and rebound conditions. Graph 2 Observed toe with sign change 42

43 During toe correction investigation (Fig.25) it was observed that the tie rod joints rotation arc during wheel travel were slightly deviated from the wheel rotation arc. Hence using the geometry, inner and outer tie rod ball joint coordinates were modified to follow the same arc of rotation as that of wheel travel. Finally desired toe change with the same sign was achieved as shown in graph 2. Graph 3 Toe adjustment geometry Graph 4 Modified toe with negligible sign change observed 43

44 Graph 5 Caster and KPI v/s wheel travel curve For a parallel wheel travel of +80mm and -60mm a maximum caster change of 2.4 was observed. It was good to control dynamic camber change and for straight line stability. It was also expected to provide good self-aligning torque during full vehicle simulations. Moreover large caster change would also lower the ride height of the vehicle and suspension will have to be stiffen causing ride discomfort but in this case it stayed within limit. For KPI, the analysis started with static value of 16. But later it was reduced to 11 to control the camber change. Though with the smaller KPI there was an increase in scrub radius but it provided a good steering feel and balanced cornering performance. Graph 6 Front track change v/s wheel travel The observed track change for the front was ±12mm for a wheel travel of +80mm to -60mm. It did not create any major effects on vehicle dynamics. Large track change would led to increase in rolling resistance and excessive tyre wear. 44

45 Graph 7 Rear track change v/s wheel travel For rear suspension there was a track change of ±3.5 mm. It was quite small and well within the expected range Steering analysis This test was performed to check vehicle kinematic behavior when it is steered. It simulates the condition of vehicle in turn or going around a corner. The target of this steering system is to have a short turning radius, good Ackerman geometry and a better response. In steering analysis the wheel is steered over the specified wheel angle or rack travel displacement from the upper to the lower bound. All the parameters are evaluated for the condition when vehicle is at very low speed and free of any lateral forces. The application of steering motion results in a wheel displacement at a specified wheel height. For performing this analysis a steering system, suspension subsystem and test rig is required. Test parameters: Suspension assembly : Final_front Output Prefix : Steering analysis Number of steps : 150 Mode of simulation : Interactive Vertical setup Mode : Wheel center Upper steer limit : 80 Lower steer limit :-80 Control mode : Absolute Steer Input : Length 45

46 Graph 8 Steer angle vs rack displacement For a rack displacement of ±80mm maximum ideal steer angle of 38 was obtained at inside wheel with the Ackerman geometry. The target for minimum turning radius was 3.65m and with the maximum steer angle, it was achieved.(graph 8). Larger steer angle forces the wheel arch to move inside and due to which pedal assembly would not be in straight line with the driver foot. But in this case ideal steer angle being less than 40º there was less space needed for wheel envelop thus full filling the criteria of large front cabin space and there was no problem of driver leg and pedal alignment. It also provided enough space to mount snow chains if needed. Graph 9 Turn radius 46

47 Graph 10 Ackerman error 4.6 Full vehicle simulation For full vehicle simulation all previously created subsystems were assembled and complete vehicle model was prepared as explained during system modeling. During this analysis it was planned to evaluate different subsystem and to observe how they influence the total vehicle dynamics. This simulation also facilitated the option of examining the influence of component modifications, change in spring and damper rates. Except the data driven analysis, it was planned to use MDI_SDI_TESTRIG which were based on driving machine. In Adams/car during full vehicle simulation, vehicle was driven by driving machine much like a test driver would do on given instructions. The driving machine steered the vehicle, applied throttle, brake and shifted gear using clutch as per the values entered. It was also possible to instruct driving machine to switch between machine control and smart driver option. 47

48 Figure 31 Driving machine function Straight line maintain test This test is part of the handling and ride analysis of the vehicle. Long driving makes driver feel tired due to continuous steering correction. It is performed to check the ability of the vehicle to drive in a straight line. This nomenclature traditionally used is just perceptual one but not from its physical definition. This test usually represents comfort but in worst case safety. It is generally assessed through the evaluation of four characteristics such as residual pull, running straight, torque steer, braking straight. The following test was performed to check for the ability to drive in straight line without any drift, sloppiness with fixed steering and constant speed. During full vehicle modeling these tests are run in the beginning to check if all the joints and connectors are attached properly and model is test worthy or not. It also gives an idea regarding the sensitivity of steering and suspension system against road irregularities or wind disturbances. 48

49 Figure 32 Vehicle straight line test set-up The test was run with constant speed and throttle. Steering input was locked. Mode of test was quasi-static straight line. During the test in the beginning Adams lock the body s fore-aft and lateral position using primitive joint. It also tried to keep the vehicle s yaw rate and lateral acceleration zero. To remove the effect of any aerodynamic drag and scrub produced by the tyre, the throttle or brake will be adjust ed to match the initial acceleration. Once the vehicle settles down, it will deactivate the primitive joint before executing the maneuver. Finally the vehicle is allowed to run according to its geometry. Test parameters a) Initial velocity : 50kmph b) Steering Input : Locked c) Throttle control : Constant d) Gear e) f) Road type Analysis type : 4 th : Flat : Quasi static Graph 11 Longitudinal vs lateral displacement For a straight line travel of more than 200m there was negligible lateral displacement. This shows vehicle has a good straight line driving behavior and in terms of handling it will be quite comfortable. 49

50 Graph 12 Front steering angle Front steering angle vs time graph shows there was no need to adjust steering during the maneuver. Little vibration is observed just after the start but that is the time when driving machine settles down the vehicle and prepare for the maneuver. Graph 13 Inclination angle vs time 50

51 Graph 14 Lateral force vs time The camber angle change was same for both set of left and right tires. Lateral forces generated in tires were cancelled being equal and opposite both in front and rear. Thus it was concluded that vehicle has a good weight distribution and symmetric set up. It is capable of driving in straight line without much steering correction and is ready for further tests. 51

52 4.7 Single Iso-lane change maneuver This maneuver is most operated handling maneuver on highways and public roads. The purpose of this maneuver is to overtake a vehicle appearing infront on a highway or to avoid running over an object/obstacle appearing suddenly. Since vehicle changes one lane it is called single lane change. While changing lane both stability and controllability are mainly evaluated by studying steering and yaw characteristics. Figure 33 Single and double lane change In this analysis, the driving machine drives the full vehicle through a lane change as specified in ISO-single lane change document. During analysis, a longitudinal controller maintains the chassis velocity as specified and later controller module acts on the steering system to maintain the vehicle on the desired lane change path. Iso_lane_change.dcd file is used to define the manuever and trace on the XY-Plane. This test was performed for three speeds. Test procedures: a) Experiment Name : ISO-Lane change b) Static setup : Straight c) Initial speed : 50, 55, 60 kmph d) Gear Position : 3 e) Step size : 0.01 f) Distance : 250 m 52

53 Graph 15 Chassis displacement This test was repeated for three different speeds 50, 55, 60 kmph respectively to evaluate vehicle yaw behavior. During simulation vehicle did change lane successfully within given parameter but at the end of maneuver it drifted slightly laterally, evident from above figure. This drift was progressive with the increase in vehicle speed. That means in real situation by the end of maneuver driver will have to do the steering correction to bring the vehicle back in straight line. Graph 16 Yaw rate and steering wheel angle at 50kmph 53

54 Graph 17 Yaw rate and steering wheel angle at 55kmph Graph 18 Yaw rate and steering wheel angle at 60kmph From the above 3 curves it was observed that at 50 Kmph vehicle s yaw rate, steer angle and lateral acceleration were in good phase but as speed increased this phase difference became evident. This difference in phase depicts that vehicle is not going in the direction where steering points, rather it is drifting away and it can be dangerous at high speed. Though the phase difference is not very large and vehicle has a top speed of 70 kmph, the extreme situation is not expected and it will be safe during the maneuver. 54

55 4.8 Ramp steer The main purpose of this test is to evaluate the understeer behavior of the vehicle while negotiating a turn. To simulate the same condition it is planned to run the vehicle in spiral path (as shown in fig.) at constant speed while changing the radius at constant rate. Vehicle entry path Figure 34 ramp steer test During this analysis, Adams/car ramps up the steering input from an initial value at a specific rate, which in this case was 5deg/sec. At the end of this test a time-domain transient response metrics is obtained. The most important quantities which are to be measured include steering wheel angle, yaw rate, vehicle speed, lateral acceleration and maneuver radius. All these quantities will be used to evaluate the steering behavior. As Adams/car does not have a default request inbuilt to evaluate understeering characteristics hence it was calculated in excel manually with the obtained data of steering wheel angle and maneuver radius. Test parameters: Figure 35 Ramp steer test parameters 55

56 Along with the above parameters a separate event was created using event builder option to keep the velocity constant and steering input changing. First simulation was done with normal conditions then the generated.xml file was edited to have the desired driver input of keeping speed constant. Figure 36 Event builder As the understeer curve is not directly calculated in Adams/car a separate excel file was generated. To calculate the steer angle at ground a separate simulation for steering system was done in static condition. Correlating factor between wheel and steering was calculated from the same. It was then multiplied with the values of steer angle of ramp steer analysis to get the mean steer angle in dynamic condition and then plotted against L/R. 56