Int. J. Vehicle Design, Vol. 52, Nos. 1/2/3/4, 2010
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Technical note: Remodelling, design and testing of a steering system for improvement of the characteristics Burak Yelken* Temsa R&D Company, Tubitak-MAM Technology Free Zone, 41470 Gebze, Kocaeli, Türkiye E-mail:
[email protected] *Corresponding author
Can Çoğun Mechanical Engineering Department, Gazi University, 06570 Maltepe, Ankara, Türkiye E-mail:
[email protected] Abstract: In the literature, mostly theoretical information and definitions on the steering system and components are found. In this study, a design map guiding the design of a complete steering system of a bus is given for design engineers. Design changes for an existing steering system are made by performing geometrical analysis, solid modelling and strength analysis of the parts. Some of the designed parts are manufactured and some of them are supplied from vendors to construct the steering system of the vehicle. The steering system is tested in the field under various testing conditions. The test results indicated an improvement in the performance of the designed steering system. Keywords: steering system; force analysis. Reference to this paper should be made as follows: Yelken, B. and Çoğun, C. (2010) ‘Technical note: Remodelling, design and testing of a steering system for improvement of the characteristics’, Int. J. Vehicle Design, Vol. 52, Nos. 1/2/3/4, pp.237–251. Biographical notes: Burak Yelken received his MSc in Mechanical Engineering from Gazi University, Mechanical Engineering Department in Ankara. He has six years of experience in the automotive industry. He is working at Temsa Global Company (Kocaeli, Turkey) as a Research and Development Project Engineer. Can Cogun is a Full Professor at Gazi University, Mechanical Engineering Department in Ankara. He holds a PhD in Production Engineering. His research and teaching concentrate on manufacturing processes, manufacturing systems, production engineering and non-traditional machining methods, with specific attention on electric discharge machining. He has published extensively in books, articles in refereed journals, conference papers in manufacturing field. He is a Consultant to industrial companies, professional organisations, and suppliers, TUBITAK (Scientific and Technological Research Council of Turkey) and the Ministry of Industry and Trade in Turkey.
Copyright © 2010 Inderscience Enterprises Ltd.
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Introduction
Steering system is one of the most important systems of the vehicle that performs the turning movement of the vehicle. It affects the dynamic and stability characteristics of the vehicle (Gillespie, 1992; Nunney, 1992). Turning movement is provided by a geometry called as Ackermann geometry (Sethi, 1991). An error in this geometry causes the slip and wear of the tyres, deterioration of the turning characteristics, insufficient turning diameter and overloading of the components of the steering elements by excessive forces and moments (SAE J695 DEC 89, 1989). The main goal of a steering system design is to obtain ideal Ackermann geometry (Heisler, 2002; Chisholm, 1984). In order to determine the components of the steering system, the geometric analysis of all components, the description of the positions of the elements during turning movement, the force and buckling analysis of the components under maximum loading conditions should be performed. In the literature, mostly the types and elements of steering system is defined (Reimpell, 2001; Birch, 1999), Ackerman geometry defined by formulas, the effects of Ackerman geometry on the vehicle and mentioning the optimum Ackerman geometry (Miller et al., 1991), different types of steering mechanisms, diagnostic steering, troubles and servicing steering system are given (Crouse and Anglin, 1976) In this study, the focus is given on the performance improvement of an existing steering system of an urban bus by performing geometrical analysis, modelling the components and force analysis. At the end of all studies, the predesign is obtained. Also, the standard field tests for testing a steering system are applied to the predesign and satisfactory results confirming the performance of the design are obtained. One of the distinguishing aspects of this study from the other theoretical works is the used practical approach in designing the steering system. Although, no optimisation approach is used in designing the steering system, the design procedure steps used in the study, namely, draft design, geometric analysis, force analysis and on-site standard tests for checking the functionality the designed steering system are very informative and useful for design engineers in the field.
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Studies conducted
In this study, the following procedure is applied in determining the steering geometry and properties of the components; 1
reperforming the geometrical analysis using the existing components of a steering system of another bus
2
improvement of the steering system geometrically by considering the wheel and the spring leafs movements
3
remodelling of the components
4
performing the kinematic analysis of the steering system by a computer
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performing the force analysis by determining the forces acting on the drag link and drop arm
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confirmation of the design by field tests.
Remodelling, design and testing of a steering system The present approach is given as follows: Determination of the steering geometry and the components – Geometrical analysis – Correction of the geometry according to turning angles
Modelling components in 3D
Force analysis – Drop arm force analysis
Force analysis – Drag link static force (compression and tension) and buckling analysis (before geometric improvement)
Force analysis – Drag link static force (compression and tension) and buckling analysis (after geometric improvement)
Force analysis – Drag link static force (compression and tension) and buckling analysis (the third situation)
Tests performed – Determination of the vehicle turning diameter – Steering moment test
Approval of steering system design
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2.1 Determination of the steering geometry and the components Some of the components of the developed steering system (steering wheel and column, steering gearbox) are available at another vehicle in use. In this study, drag link, drop arm and the other components on the front axle are redesigned and manufactured. The components, which are on top of the front axle are supplied from the manufacturer in an assembled and adjusted form. The correct toe-in and toe-out adjustments are essential for a stable driving and good dynamic behaviour of the vehicle. Initially, the front axle CAD model is created. The front axle is positioned on the model according to the vehicle axle distance information. The position and the angle of the steering wheel are determined by the ergonomic considerations. The position of steering box, which is assembled to the steering column and steering wheel, is selected temporarily. According to this position, the CAD model of the steering box is created and added to the CAD assembly model of the vehicle. The initial length of the drop arm and drag link are determined by performing geometric improvement analysis. For this purpose, CATIA Part Design and Sketch Modules are used (CADEM, 2005). As a start value, the vertical position angle of the drop arm is taken as 90°. The draft sketch is drawn and the path of the drag link is determined graphically (Figure 1). Information about available lengths of the drop arms is provided from the manufacturer company and the closest length to the steering arm is chosen as drop arm for the better vehicle stability. Drag link length is determined at the end of this study. At the beginning, it was assumed as an approximate value. The displacement of the leaf springs in use is 100 mm downwards and upwards. For both cases, the positions of leaf spring and the drag link are determined (Figure 2). At the upper position, there is 1 mm deviation (Figure 3) and at lower position 10 mm deviation from the ideal position (Figure 3) considering the connection point between steering arm and the drag link. The deviations are found acceptable when the magnitudes of the deviations are compared with the length of steering arm (225 mm). Figure 1
Movement path of drag link
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Figure 2
Sketch of the steering system when the leaf spring is at down position
Figure 3
The top view of steering arm, (a) upper position (b) lower position (see online version for colours)
(a)
(b)
By using these deviations, the angular deviations are also calculated. Steering arm angular deviations at upper position is; ⎛ 1 ⎞ Arc tan ⎜ ⎟ = 0.25° ⎝ 225 ⎠
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Steering arm angular deviation at lower position is: ⎛ 10 ⎞ Arc tan ⎜ ⎟ = 2.54° ⎝ 225 ⎠
According to literature search and past experiences, the upper and lower position deviations should not exceed 0.5° and 3°, respectively. If the deviations are not in the acceptable limits, the length of the drop arm will be changed until acceptable values are obtained. In the second stage of calculations, rotation angles of drop arm are checked at the lower and upper positions. The fix point of drop arm is determined in the second stage of calculations. The approximate position of the steering gear box was determined according to ergonomic considerations. Accordingly, the position of the drop arm which is assembled on steering gear box is also determined. The vertical angle of drop arm was assumed 90° at the beginning. The turning angles of the vehicle are fixed for the front axle as 50° to right and 35.2° to left (Figure 4) by the supplier company. Figure 4
Steering turning angles
During the turning motion of vehicle, the wheel turns over the kingpin axis. On this axis, steering arm and ball joint are determined and the paths (forward direction 50° and backward direction 35.2°) are drawn. At the beginning of the analysis, drop arm was assumed at the vertical position. The position of the drag arm is not found in bisector position according to the projection points of steering arm movement limit points on drop arm movement circular path. In this case, the adjustable parameters are the angle of the drop arm and the length of the drag link. The angle of the drop arm is changed for being bisector according to the critical angles provided by the supplier. By this way, the number of steering wheel rotations to right and left become equal. Before the modelling of the components, calculations are repeated for the last position of the drop arm in order to be sure that the deviations are in the range of acceptable limits (Figure 5).
Remodelling, design and testing of a steering system Figure 5
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The position of drop arm at the end of the geometric analysis (see online version for colours)
2.2 The modelling of the steering system components After calculations and confirmations, front axle, steering gear box, drop arm and drag link and axle connections are modelled by CATIA software according to the values provided by the suppliers (CADEM, 2005). The models of the components are used in force analysis in later stages (Figure 6). After the modelling process, the kinematic analysis module of the CATIA software is used to check the possible contacts between components when the wheel is rotated around kingpin axis 50° forward and 35.2° backward. Figure 6
The modelling of the leaf springs (see online version for colours)
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Force analysis
3.1 Drop arm force analysis The force analysis of steering system is performed for drop arm and drag link, which are exposed to the moments and forces generated by steering gear arm at maximum steering hydraulic pressure (maximum moment and force). Forces and stresses are modelled by using finite elements method with MSC Patran program and solved by MSC Nastran program. The moment (so the applied force F), which is acting on drop arm by steering gear box is calculated and stress distribution on it are found. Drop arm is assembled on steering gear box. It transmits the moment generated from steering gear box in force form to drag link by drop arm (Figure 7). According to the information provided by the supplier, the maximum pressure the pump can generate is 130 bar and at this level of pressure, the rotational moment on drop arm – exerted by steering gear box – is 2093 Nm. Figure 7
Drop arm-drag link force transmission (see online version for colours)
The force F acting on the drop arm at drag link assembly point is;
F=
2093 Nm = 10733.3 N 0.195 m
The material of drop arm is forged steel 40NiCrMo4. Some of the mechanical properties are as follows: elasticity module
210,000 MPa
Poisson ratio
0.29
density
7,700 kg/m3
yield strength
931 MPa.
10733.3 N force is acting vertically on drag link connection. According to the analysis, the maximum stress at the connection point of drag link and drop arm is 650 MPa. Since the yield strength of the material is 931 MPa, there is no any strength problem for the proper functioning of drop arm.
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3.2 Drag link force analysis For the drag link, static force analysis under compressive and tensile load and buckling analysis under compressive load are done. The material of drag link is steel (40 × 5.7 mm pipe) and some of the mechanical properties are as follows: elasticity module
210,000 MPa
Poisson ratio
0.3
density
7,800 kg/m3
yield strength
355 MPa.
The component is meshed into finite elements before the analysis. Mesh type is ‘tetrahedral solid element’. 46,811 elements are formed with 12,040 nodes. The material used in the analysis is assumed to be ‘linear-elastic’. The boundary conditions used in the analysis are: a
X axis passes through the sphere centres of both ends of drag link
b
the coordinate system is located on one of the sphere centres
c
the sphere at the steering arm connection have linear degree of freedom in x, y and z axis and rotational degree of freedom around x axis
d
other sphere centre has only linear degree of freedom in x, y axis and force applied in y direction.
Force analysis of the first situation (before geometric improvement) As the first situation – before the geometric improvement – the angle between drop arm and drag link is determined by using CATIA software and the force ( F ′) applied on drag link by drop arm is calculated (Figure 8). Figure 8
The F ′ force applied on drag link for the first situation (see online version for colours)
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The calculated compressive force is applied to the element (Figure 9) using Patran program and solved according to boundary conditions using Nastran program. At the end of the analysis, it is found that maximum Von Mises stress is 140 MPa (Figure 9). This value is less than the yield strength of the material (355 MPa). The meshing is the same for tensile force analysis. Only, the direction of the applied force is reversed. The maximum Von Mises stress value is 140 MPa and this value is less than the yield strength of the material. Figure 9
Static force analysis at the first situation (compressive force) (see online version for colours)
Figure 10 Buckling analysis for the first situation (see online version for colours)
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Linear elastic buckling analysis is performed by spherical joint conditions using the same meshing introduced in the previous cases. Results indicate that for 18.2 kN compressive force the factor of safety is 2.78. According to the static analysis, the stresses are less than the yield strength of the material (Figure 10).
Force analysis of the second situation (after geometric improvement) According to the findings in this study, a drag link is manufactured and tried on the vehicle. But, it is found that it is very difficult to rotate the steering wheel. As the second ′ ) applied to drag link by situation – with the geometrical improvement – the force ( Fopt drop arm is increased (Figure 11). ′ force transmitted to the drag link (see online version for colours) Figure 11 Fopt
′ is calculated as; At this position (Figure 15), Fopt
′ = Fopt
10733.3 = 25783.8 N Cos 65.6
As a result of static analysis, it is found that maximum Von Mises stress value on the drag link is 199 MPa. This value is less than the yield strength of the material. For static force analysis with tensile load, the meshing is the same as before, except that the direction of the applied force is reversed. The maximum Von Mises stress value on the drag link is 199 MPa. This value is less than the yield strength of the material. The buckling factor of safety of the material for 25.8 kN compressive force is found as 1.25. This value is risky for a steering system.
The third situation of force analysis In order to eliminate the risk of low factor of safety in the second situation, a 42 × 6.7 mm pipe is used instead of 40 × 5.7 mm pipe. The applied force is the same as ′ = 25, 783.8 N. According to the force analysis with the second situation and it is Fopt compressive force on the drag link, maximum Von Mises stress value found as 117 MPa. This value is less than the yield strength of the material. For pulling force analysis, meshing is the same as before. Only the direction of the applied force is reversed. The maximum Von Mises stress value on the drag link is 117 MPa. This value is less than the yield strength of the material. Linear elastic buckling
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analysis results show that for 25.8 kN compressive force the factor of safety is 3.24. According to the static force analysis, the stresses are less than the material yield strength. So, there is no plastic deformation in the material.
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Tests performed
In the first situation (before the geometric improvement), the steering moment measurements could not be taken since the system was disassembled in a short while. The problem of ‘excessive moment requirement to turn the steering wheel’ is reported by all test drivers. Also, after testing the vehicle for some time, a part of the drag link is found to be worn due to the contact with the wheel (Figure 12). Figure 12 Wear of the drag link (see online version for colours)
After the geometric analysis and improvement studies of the steering system, the vehicle is tested by the test drivers. They reported the reduced steering moment and better steering conditions. Later, the steering system is tested in the field at various driving conditions. The following tests are planned and applied to the vehicle: 1
Determination of the vehicle turning diameter: Since the designed vehicle is an urban bus and is driven in the city, the manoeuvring ability is very important. SAE J695 Standard of Turning Ability of the Vehicles (Robert Bosch GmbH, 1993) is used to determine the turning diameter of the vehicle. First of all, all systems are checked and the vehicle is loaded to gross weight of 12 tons. The vehicle is driven in the factory open field at low speed (10 km/hr) by turning fully the steering wheel in both sides and then the traces of the wheel on the ground (wheels are wet) are observed. In measuring the turning diameter from pavement to pavement, a full circle of wheel track is obtained. The track of wheel from outside is measured as 7.73 m to the left and 8.06 m to the right side. In measuring the turning diameter from wall to wall, a half circle track is obtained by starting and ending the turning movement by lines on the ground which are the ground projections of side walls of the vehicle. The measured values are 8.97 m to the left and 9.22m to the right. The measurements are compared with the vehicles in the same segment and are found acceptable.
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Steering moment test: Steering moment is measured in two different ways: a According to 70/311/EEC European Council Directives, steering force should not exceed 25 kg while the vehicle turns on a 12 m diameter circular path (70/311/EEC, 1985). For this test, a 12 m diameter path is drawn on the ground and a torque meter is assembled between the steering wheel and the steering column for measurement. When driving, the speed of the vehicle is constant (10 km/hr). The measurements are taken after four seconds and totally, 500 data are collected and recorded by the connected data logger. The procedure is repeated for both left and right directions. During the tests, the maximum recorded value is 15.1 Nm (Figure 13) for left turn and 18.8 Nm for the right turn. Since both values are less then 25 kg, the designed steering system is compatible with the rules set by ‘European Council Directives – Steering System Components of Motor Vehicles’ (70/311/EEC, 1985).
Figure 13 Steering moment-time graphics for left turn (see online version for colours)
b
3
The other steering force test is performed according to Steering System Loads and Force Test Standard (Rover Engineering Standard – Steering System Load and Force Test Standard) (RES.61.11.500, 1980). In the test, all steering and suspension system components are checked and then the vehicle is loaded to 13.5 tons, which is the permissible weight. A test route of 250 km is decided including urban, intercity and highway drivings. Test lasted for five days and two test drivers are used during the test period. The maximum steering torque values measured during the tests are 19 Nm to the right turn and 15.2 Nm to the left turn.
Pavement hit test: This test is performed according to Steering System Loads and Force Test Standard (Rover Engineering Standard – Steering System Load and Force Test Standard) (RES.61.11.500, 1980). Wheel pressures are checked and the vehicle
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B. Yelken and C. Çoğun is loaded to 13.5 tons. A pavement with 115 mm height is built to the test area. The front-left wheel of the vehicle is hit to the pavement with 20±5 km/hr speed and with 30–35° hit angle. At the instant of hit, the brakes are not applied. Test is repeated 20 times. The test is also repeated for the front-right wheel. After the test, all steering system components, especially steering gear box, drag link and drop arm, are carefully checked for bending, distortion and loosening of the connections. No plastic deformation is observed on the drag link and other system components.
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Discussion
While considering the results obtained from geometric analysis, calculations, design of the components and tests performed the followings should be kept in mind: 1
The force analysis performed for drop arm and drag link are static type (tension, compression and buckling). The analysis of dynamic stresses on steering system components is not performed.
2
The theoretical and experimental fatigue analysis of the components of the steering mechanism is not performed.
3
The introduced study to improve the steering system characteristics and components is not an optimisation study. The aim of the study is to reach the conditions which hold the predefined acceptable values.
4
The drag link is continuously exposed to compression and tension type of loadings. So, the stress analysis of this component is performed for both types of loadings.
5
Although a detailed literature search is conducted to learn more about the factor of safety of the stress values of the designed drag link and drop arm components, no solid information is found. The drop arm suppliers advice of 1.4 (acceptable factor of safety for the stresses) is taken as reference in design calculations.
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Conclusions
The main aim of this study is to introduce the improvement of an existing steering system by applying different geometrical and mathematical analysis, and test respects. It is concluded that geometrical analysis of the steering system must be performed using modelling software before assembling the components of steering system for the field tests. It is also shown that at the design stage of drag link, the forces acting on steering wheel box, drag link and steering arm should be determined and the analysis of force and buckling must be performed. In the first situation, no geometric improvement efforts are put forward on the existing steering system. In this case, although no problems concerning the static forces and buckling are observed, it is found that it is very difficult to turn the steering wheel. Also, it is observed that the wheel rubs to the drag link causing wear. In the second situation, improvements (design changes) are made in the steering system. Although the force on the new drag link increases slightly, the buckling analysis of the material gives low factor of safety. In the third situation, the thickness of the material used for drag link
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is increased to prevent the buckling risk of the component. Acceptable results are obtained from the force analysis for compressive and tensile loadings and buckling for the modified drag link. The manufactured steering system is tested under various field conditions. It is found that the new steering system holds the standards and procedures described for the field conditions. The newly designed and tested steering system is now in use.
References 70/311/EEC (1985) European Council Directives – Steering System Components of Motor Vehicles, Society of Automotive Engineers, pp.2–21. Birch, T.W. (1999) Automotive Suspension and Steering Systems, Delmar Publishers, Albany, pp.237–269. CADEM (2005) CATIA V5 Learning Book, Infrastructure, pp.58–79. Chisholm, I. (1984) Automobile Engine and Vehicle Technology, McGraw-Hill, New York, pp.252–286. Crouse, W.H. and Anglin, D.L. (1976) Automotive Chassis and Body, McGraw-Hill, New York, pp.55–125 Gillespie, T.D. (1992) Fundamentals of Vehicle Dynamics, Society of Automotive Engineers , Warrendale, PA, pp.275–307. Heisler, H. (2002) Advanced Vehicle Technology, Butterworth-Heinemann, Oxford, pp.311–367. Miller, G., Reed, R. and Wheeler, F. (1991) ‘Optimum Ackerman for improved steering axle tire wear on trucks’, SAE Paper, 912693 Nunney, M.J. (1992) Light and Heavy Vehicle Technology, Butterworth-Heinemann, pp.367–412. Reimpell, J. (2001) The Automotive Chassis, Butterworth-Heinemann, Oxford, pp.266–306. Robert Bosch GmbH (1993) Automotive Handbook Bosch, Wiley, Stuttgart, pp.101–145. Rover Engineering Standard – RES.61.11.500 (1980) Steering System Load and Force Test Standard, Rover Engineering, pp.1–17. SAE J695 DEC 89 (1989) Turning Ability and Off Tracking-Motor Vehicles, SAE International Publishing, Warrendale, pp.5–15. Sethi, H.M. (1991) Automotive Technology, Hill Publishing Company Limited, New Delhi, pp.112–134.
