Design of a Low Power Active Truck Cab Suspension

Transcription

1 Design of a Low Power Active Truck Cab Suspension João Vitor Régis Sampaio DCT Nº: Coaches: prof.dr. Henk Nijmeijer dr. ir. Igo Besselink ir. Willem-Jan Evers Eindhoven University of Technology Department of Mechanical Engineering Dynamics and Control Eindhoven, November 30 th, 2009

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3 Table of Contents Chapter 1: Introduction Background Delft Active Suspension Other researches using the DAS principle Active truck cab suspension Problem statement and objectives Outline... 5 Chapter 2: Analysis of the basic DAS concept Steady state model Results Conclusion... 8 Chapter 3: Proposed solutions Semi-circle Mathematical model for the semi-circle mechanism Motor force variation and effective stiffness of the system Results from steady state analysis Realization and feasibility Balance bar Mathematical model for the balance bar mechanism Motor force variation and effective stiffness of the system Results from steady state analysis Realization and feasibility Wishbone length variation Mathematical model for the wishbone length variation Realization and feasibility Tetragon geometry Results from steady state analysis Continuous variable transmission Conclusion i

4 TABLE OF CONTENTS Chapter 4: Parameter analysis of the selected designs Study of the semi-circle geometry Study of the Tetragon geometry Conclusion Chapter 5: Mechanical design Specification of the available space Study of the DAF CF Study of the DAF XF Actuator s elements Spring specifications Solutions for the front cab suspension for the CF Selection of the motor Solutions for the front cab suspension for the XF Solutions for the rear cab suspension Conclusion Chapter 6: Conclusion Recommendation for further work References ii

5 Chapter 1 Introduction The main function of a suspension system is to reduce the effect of environmental vibrations to the suspended system, e. g. to limit the transmissibility. In a truck, the comfort of the driver is essential. As commented in references [1], [2] and [3] The backache has the second place in the common health problems of the truck drivers (just behind the sleep disordered breathing). The work posture, the long time in the same posture, the vibration of the tractor during the day are some of the clues. These issues make an active cabin suspension worthwhile to research. In trucks, active cabin suspensions can be used to improve the drive comfort, but usually the energy consumption is high. As a low power actuator is desirable, a new design is needed. The focus in truck cab suspension development has the benefit of not deteriorating the truck s handling behavior. Furthermore, the effort to control this secondary suspension requires less energy, because the suspended mass supported by such system is lower than the total mass supported by the primary suspension. The strategy to attend the low power requirement is based on the Delft Active Suspension (DAS) principle [4], which seems to be very promising as it has a relatively low energy consumption and good performance. However, the current design has some draw-backs with respect to packaging, non-linear stiffness and durability issues. Within this report, is detailed the steady state analysis of multiple variable geometry mechanisms based on the DAS principle, which can be applied in suspension systems. Furthermore, the design steps of a single selected solution are described with the intent to implement it in a truck. The main contribution of this report is an overview of various 2D variable geometry solutions, based on DAS principle, which can be used in an active cabin suspension. Furthermore, a single design is detailed to the specifications of a real truck Background In passenger cars, the insertion of an actuator between the sprung and unsprung masses can be used to minimize road induced vibrations. Furthermore, the actuator can also introduce an external force in a suspension system to compensate the lateral load transfer during cornering, suppressing the roll movement of the vehicle [3]. The difficulty in obtaining a mechanism which works at high force variations is related to the high power consumption involved in the quick response of the actuators. Nowadays, active suspensions Figure-1: Illustration of a full bandwidth active suspension system [5].

6 CHAPTER 1- INTRODUCTION often use an electro-hydraulic system to control the vehicle attitude (see Figure-1). The amplitude order of the forces that act on the system is large and there is also a significant weight penalty, compared with passive suspension. This is the result of all the components that are necessary for the operation of this system (pump, reservoir, pipework, filter cooler and etc) [5] Delft Active Suspension In the 1990 s, the Delft University of Technology developed the so called Delft Active Suspension (DAS), a mechanism to vary the suspension force by changing the geometry of the mechanism (sometimes called as variable geometry active suspension). The actuator force is adjusted by the suspension lever ratio / (see Figure-2a). The great difference, comparing with other active suspension systems, consists in the transversal work of the actuators in relation to the spring course. The energy needed to control the lever ratio variation is much lower than the energy involved in the increment of the spring force. Thus, the main improvement of this new concept is the low power consumption, since is not the spring force that is been controlled (the spring elongation can even remains unchanged). The feasibility requirements led them to create a 3D rotary geometry. The so called conemechanism (see Figure-2b) follows the same lever ratio variation principle. Moving one end of the spring around the base of a cone, while the other end is at the vertex, should give the idea of non-working and no energy consumption, since the spring length remains the same. Figure-2: (a) Variable compensating force with adjustable mechanism (left), and (b) the Delft Active suspension (right) [4]. Even with this geometry, power consumption can be substantial when the mechanism is not in the ideal conditions (when the suspension displacement is not zero). Furthermore, frictional resistances at the joints also affect the energy consumption. The DAS was built and implemented in a passenger car to tests and evaluate the performance of this system. As result, it presented good response to control the angular displacement of the vehicle (maximum peak of 1º - see Figure-3), and also a lower power consumption comparing with other active suspension concepts (average of 770 Watt and peak of 3kW in a severe double lane change) [4]. Figure-3: Double lane change at 95 km/h (ISO/TR 3888) [4]. In the double lane change test, the vehicle lateral acceleration changes direction more than five times in less than six seconds, what makes the electric motors accelerate with the maximum possible 2

7 CHAPTER 1- INTRODUCTION speed to control the vehicle attitude. This is the main cause for the peaks in power consumption in combination with losses in electric system and friction losses. The Delft Active Suspension is a promising concept for a low power attitude control system, yet the modifications needed to implement this system in a test vehicle were considerable. Hence, it is necessary to look for more feasible solutions Other researches using the DAS principle Early studies on the development of low power active suspension are reported in [9]. Its goal is to present and analyze 2D geometries based on DAS. Some of those mechanisms are represented in Figure-4. Figure-4: Others solutions for the DAS principle [9]. The benefit of these ideas is mainly the flat geometry, which reduces the occupied space. However, the reliability issues still have to be analyzed. Some of those ideas are not feasible for application in a truck cabin suspension. The use of rollers and sliders are not desirable, because there is a high friction associated with the work of those components as well as high wear. Thus, the development of a low-power cabin suspension is not finished. This study just focused on steady-state analysis of different mechanisms, evaluating the behavior for the stiffness and forces variation without advance in constructive issues. However, specifications about the system weight and dimensions still need to be analyzed to complete the feasibility study. Furthermore, another important challenge is to fit such device in a truck cabin. Another previous design for passenger cars focused on a 3D geometry [12]. The electromechanic Low Power Active Suspension (elpas) is presented in Figure-5. This device consists of a wishbone connected to a frame. An electric motor is attached to the wishbone, and it controls the position of a rotational arm connected to a bar. This bar transfers the load for the spring. Figure-5: (left) Diagram of the elpas geometry. (right) The elpas mechanism [12]. 3

8 CHAPTER 1- INTRODUCTION The elpas presented good results for its behavior. However its mechanism follows a 3D design. In the truck cabin, the space available for the suspension system is limited. That is the reason a packaging solution is desirable. Furthermore, the arm controlled by the motor works under high bending moments condition Active truck cabin suspension The idea of reducing the vibrations through the utilization of an active suspension system seems interesting for application in a truck cab suspension. Separating the system in two suspension levels eliminates the conflict between handling behavior and driving comfort. The cabin suspension can be specifically designed to accomplish the comfort requirements, while the primary suspension can be designed with focus on handling. Furthermore, using this configuration it is possible to work with lower forces magnitude, since the suspended mass supported by the cab suspension is lower than the mass supported by the primary suspension. In [7], a strategy of self-powered active suspension is presented. In this study, it uses the energy dissipated in the primary suspension work to control the active cabin suspension. It concludes that an enough amount of energy is associated with the energy absorption of the primary suspension, which used to be completely dissipated by dampers. Thus, the strategy consists in use this amount of energy to control the cabin suspension system. To accomplish this aim, they proposed the use of a regenerative damper which stores the vibration energy in a condenser (see Figure-6). Figure-6: Half car model of a heavy truck with self-powered active control [7]. The solution for the cabin suspension actuator is an electromagnetic damper, composed by a DC motor, a ball screw and a nut. In this configuration, rotary motion is converted in linear motion by the ball screw and the nut (see Figure-7). Figure-7: Configuration of an electromagnetic damper [7]. Such device suggests a low power suspension scheme; however, the actuator still fights directly against the whole suspended weight. This strategy aims on the energy obtainment for the cabin suspension control system. However, the energy consumption of the cabin suspension was not reduced. This device showed good performance in comparison with semi-active and passive systems, but the 4

9 CHAPTER 1- INTRODUCTION power consumption and its behavior could be even better with the integration of a DAS mechanism in the cabin suspension Problem statement and objectives The objective of this study is to design a low power force actuator suitable for an active truck cabin suspension. The design should follow the DAS principle to reduce its power consumption. Previous designs for truck cabin focused on a 3D geometry [8]. However, for packaging reasons, a 2D geometry is desirable (cabin suspensions typically have a limited working space). Other requirements that such system should accomplish are: 1. The mechanism must be flat and small. This means that the mechanism should be fitted in the space available for the actual cab suspension system. More specifications about that are described in Chapter The total weight should be below 50 kg; 3. The actuation force must be within the range: N. Based on earlier simulation studies. 4. The effective stiffness (explained in Chapter 2) must be within the range: N/mm; 5. The maximum actuation force variation must be at least: N/s. This is necessary to respond quickly to the force variation; 6. It has to be feasible for production and low cost. This means that the mechanism proposed needs to be fitted in a truck cab suspension with a minimum of changes in the original vehicle (this is evaluated by comparing the proposed solutions) Outline This report is structured as follow. In chapter 2, the basic concept of the variable geometry mechanism is presented. Herein, the mathematic model for a steady-state analysis of such geometry is described. Results for the actuator force, effective stiffness and the motor force increment are presented as well. In chapter 3, several designs are presented and analyzed. Some particular characteristics are commented to show the advantages and the disadvantages of each solution. Furthermore, CAD sketches for some mechanisms illustrate how the solutions can be realized. In chapter 4, two mechanisms are selected for a more detailed design study. The aim here is to observe which parameters can control their behavior and find the best configuration to their geometries to accomplish the requirements. In chapter 5, the space available to fit the cabin suspension is specified. Furthermore, some solutions for implementation of the active suspension system on the truck model XF and CF are presented. Criteria for the selection of the mechanical components (electric motor and spring) are described as well. In chapter 6 ends with conclusions and recommendations. 5

10 Chapter 2 Analysis of the basic DAS concept An initial mathematic model for a 2D mechanism is made in order to analyze the magnitude of the forces involved. As showed in the picture below, the mechanism works ranging the lever ratio by changing the spring position. The system has two degrees of freedom (α and x), where α represents the angular position of the wishbone and x represents the spring position along the wishbone (see Figure-8). The aim of this simple model of the secondary suspension system is to evaluate the following parameters: The force that effectively acts on the suspended mass; The increase of the motor force when the system works out of ideal conditions ( 0), In a steady state condition, this force is needed to keep the spring position; The effective stiffness of the suspension system at the end of the wishbone. It represents the variation of the actuation force, keeping the spring in a fixed position, and for a small angular displacement of the wishbone (around: 0) Steady state model Figure-8: Variable geometry active suspension concept Under steady state conditions 0, 0, the actuator force is given by:.. (2.1)

11 CHAPTER 2- ANALYSIS OF THE BASIC CONCEPT Herein, the spring force is given by:.. sin, (2.2) With being the spring stiffness and its preload. Combining (2.1) and (2.2) gives:... sin, (2.3) The deflection of the wishbone gives causes a motor force to appear, which is needed to keep the spring in place:. sin. sin.. sin, (2.4) Furthermore, the effective stiffness at the end of the wishbone is given by: Changing the variable, as For this analysis, the following data given in Table-1 is used:, (2.5)., (2.6) Table-1 - Parameters used. Parameter Value Unit Meaning L 150 mm wishbone lenght F act N nominal actuator force (1/4 cabin weight) C s - steel 60 N/mm reference stiffness for steel spring C s - air 25 N/mm reference stiffness for isolated air spring The travel course for the spring position was standardized for both steel and air spring. Hence, the total travel course equal to the wishbone length is adopted ( 150 mm). Thus, using the force range specified by requirements, the spring preload is given by: Where it is found: = 6400 N.. _., (2.7) 2.2. Results Using the parameters and equations of the previous section, it is possible to plot the results. In Figure-9, the actuator force in relation to the spring position along the wishbone can be seen for two different spring stiffness. Figure-9: Actuator force vs. spring position along the wishbone for some values of alpha. With a steel spring (left). With an air spring (right). 7

12 CHAPTER 2- ANALYSIS OF THE BASIC CONCEPT In this situation, it is possible to observe that the magnitudes of the forces involved are almost the same. The magnitude variation is larger with the steel spring because it has a higher stiffness than the air spring. In Figure-10, the motor force variation in relation to the spring position along the wishbone can be seen for two different spring stiffness. Figure-10: Motor force increment vs. spring position along the wishbone for some values of alpha. With a steel spring (left). With an air spring (right). It is natural to conclude that the spring stiffness influences the magnitude of the motor force variation. Since the motor force is one of the parameters that influences the energy consumption, it is desirable to reduce the magnitude of this parameter. A lower value for the spring stiffness can attend such aim. In Figure-11, the effective stiffness of this mechanism in relation to the spring position along the wishbone can be seen. The effective stiffness for the both springs is not within the range: 10 < Ceff < 30 N/mm, and is a non-linear function of. Although, it is observed that an intermediate value for the spring stiffness could both reduce the forces magnitude variation on the system and attend the effective stiffness requirement if necessary Conclusion Figure-11: Spring stiffness variation vs. spring position along the wishbone. This is a very simple geometry, but the use of sliders in each extremity of the spring makes it not reliable (in such kind of union great wear can be observed, and debris accumulation can turn it even worse). Furthermore, this device has a moving spring, which is a not desirable characteristic as the acceleration of such moving mass will influence the energy consumption. The analysis of the basic concept gives an introduction to the problem, showing what requirements are difficult to accomplish. Furthermore, it shows how the elements interact to solve the problem. 8

13 Chapter 3 Proposed solutions The basic concept, presented in chapter 2, does not attend the requirements proposed, mainly because of its feasibility. Such mechanism has joints that turn the device not reliable (sliders); furthermore, the effective stiffness has a non-linear format that does not fit the requirements. However, the analysis of the basic concept gave an introduction to the problem, showing what requirements are difficult to accomplish. Furthermore, it shows how the elements interact to solve the problem. In this chapter, the analyses of more possibilities of solutions are presented. It is the summary of a brainstorm process, in which ideas are described and analyzed. Thus, for each solution, a short description of how the mechanism works is given as well as the mathematic model that governs its steady-state behavior. Furthermore, a CAD scheme for the mechanism is shown to observe how it could be built and what kind of mechanical elements would be involved. Again, the actuator force, the motor force needed to maintain the spring position, and the effective stiffness are plotted to evaluate the geometries. The mathematical results show if the solution has the possibility to accomplish the requirements. Herein, the CAD scheme complements such analysis, presenting a first approach of feasibility and reliability evaluation Semi-circle The Semi-circle mechanism, obtained from reference [9], consists of circular motion of the spring force along a circular path referenced in the wishbone. The center of such arc is the point of spring fixation at the chassis (for: =0). Thus, the variation of the spring force momentum in relation to point O results in the variation of the actuation force magnitude. In Figure-12, the parameters and are the degrees of freedom of this system. They represent, respectively, the angular displacement of the wishbone and the angular motion of the spring. Figure-12: Diagram of the semi-circle mechanism, [9] Mathematical model for the semi-circle mechanism For the evaluation of the characteristics of this design, a mathematical expression for this mechanism is presented in terms of the initial parameters. These parameters are given in Figure-12 (1). The horizontal distance from point O up to point P.

14 CHAPTER 3- PROPOSED SOLUTIONS h The vertical distance from point O up to point P. The spring length for initial configuration of the mechanism (the spring has some preload in this configuration). The wishbone length. For this purpose, the mechanism movement is decomposed in two parts (see Figure-13). At first, is introduced an angular displacement for the spring, and then is introduced an angular displacement for the wishbone. For convenience positive values are used for both parameters. Figure-13: (1) Semi-circle mechanism at natural position. (2) With an angular displacement for spring position. (3) With an angular displacement for the spring and wishbone. (4) Sketch with the relevant dimensions for the final configuration of the geometry. Under steady-state conditions ( =0; =0), the balance of momentum around point O is given by:..cos( )=., (3.1) The distance is the arm of the spring force around the point O. Herein, the spring force ( ) is given by: = +.( ), (3.2) Herein, is the spring stiffness, and is the preload. Expressions for the distances:, and in terms of initial parameters are necessary. The equation for is given by: = +h, (3.3) 10

15 CHAPTER 3- PROPOSED SOLUTIONS To describe the distance, the position of the point Q is necessary (the joint between the spring and the wishbone). It will be used the coordinate (, ) to reference this position, with the point O = (0,0) representing the origin of the system. In Figure-12(a) it is possible to observe that the values for this coordinate at the initial position are given by: = ; =h ; (3.4) Introducing an angular displacement for the spring (see Figure-12(b)), the coordinate of the point Q changes to (, ), given by: Thus, the value for the distance is given by: = +. ( ); =h. ( ); (3.5) = + ; (3.6) The coordinates of the point Q can also be represented in the circular system by (, ), with being the angular position of such point (in comparison with the horizontal line starting on point O ). This parameter is given by: =, (3.7) Even with the introduction of an angular displacement for the wishbone, the distance remains the same (see Figure-12(c)). For this situation the spring length becomes and the value for the angle determined by the wishbone and the side is modified as well. Thus, using the cosine law in the angle of Figure-12(d) it is possible to determine the final spring length ( ): = ( ) ; (3.8) With being the initial angle between the wishbone line and the side, given by: = ; (3.9) For the distance the trigonometric relation in the triangle of Figure-12(d) gives: =.( )=. =. ; (3.10) = + ; (3.11) It is not convenient to join the equations, because the mathematical expression would become too large Motor Force Variation and effective stiffness of the system The angular deflection of the wishbone gives rise to a motor force, which is needed to keep the spring in place: =. ( ); (3.12) With being the angle between the actual spring line and the ideal spring line (line of spring force actuation in ideal condition ( =0). The value of this angle can be found using the cosine law in the two triangle represented in Figure-14 ( and ) what gives: = 2. 1 ( ) ; (3.13) = ( ) =.. ; (3.14) 11

16 CHAPTER 3- PROPOSED SOLUTIONS Figure-14: Sketch of the Semi-circle geometry used to derivate the Fmotor equations. Furthermore, the effective stiffness in the end of the wishbone is given by: = ( ) (. ) ; (3.15) However, once the expression found for the is extremely large, it would be difficult to find an equation for effective stiffness by hand. Using computational numerical methods it is possible to derivate this equation. However, since a small displacement value for is used (in turn of the origin reference: =0 ), it is also possible to use the approximation given by: = ( ) (.. ), (3.16) The low expression (.sin.sin ) represents the small vertical displacement at the end of the wishbone. For convenience, a total travel of 1mm is adopted for such expression, so, the final equation for the effective stiffness become: = ( ), (3.17) With the value for the angular displacements equal to: = =0, Results from steady state analysis Implementation of the equations determined in the previous section in Matlab show the results for, and (see Figures-15 and 16). The values for the initial parameters are: =75mm, h=300mm, =400mm and =150mm. Figure-15: Analyses for the actuator force (right) and motor force increment (left) of the semi-circle mechanism. 12

17 CHAPTER 3- PROPOSED SOLUTIONS Figure-16: Analyses for the effective stiffness of the semi-circle mechanism. However, these results do not attend all requirements. They are very close to the results found in the previous chapter for the basic concept. It is possible to observe that the effective stiffness presents high amplitude of variation (from zero up to 70 N/mm). However, the increment in the motor force is lower than the basic concept Realization and feasibility A CAD scheme for the Semi-circle geometry is presented to illustrate how this mechanism could be built. For this device, it seems to be necessary the use of slider for the joint between the spring and the wishbone (point Q ); however, it is the main problem of the basic concept, so, a modification in the mechanism is proposed. Thus, it is included one more link bar to change such kind of union for rotational joints. In Figure-17, are represented the wishbone, the spring, the electric motor and the link bar that connects the spring to the wishbone. Figure-17: CAD sketch of the Semi-circle mechanism. This solution has advantages in terms of feasibility. The components are not complex to build. Such system also seems to suffer less influence of friction, as the friction force moment is lower in rotational joints. However, the position of the electric motor increases the moment of inertia of the system. In Figure-18, other configurations for the motor position are presented. Figure-18: Different configurations for the motor position at the semi-circle mechanism. (left) crank connecting-rod; (right) ball screw actuator. The location of the spring is not good either. As defined in the requirements, the actuator force variation should be at least =10000 N/s. To accomplish that, the spring should be accelerated quickly to another position, so the inertia effects of the spring mass become relevant. 13

18 CHAPTER 3- PROPOSED SOLUTIONS 3.2. Balance bar The balance bar mechanism, obtained from reference [9], consists in a lever connected to the spring at one end and also indirectly connected to the wishbone through another link. The lever ratio is controlled by the pivot horizontal position (where the motor force acts). In this case, the spring is fixed and the degrees of freedom are: the horizontal position of the pivot along the balance bar ( ) and the wishbone angular displacement ( ) (see Figure-19). Figure-19: Diagram of the balance bar geometry [9]; Mathematical model for the balance bar mechanism For the evaluation of such solution is presented a mathematical expression for this mechanism in terms of the initial parameters: The horizontal distance from point O up to point A. The horizontal distance from point O up to point B. The wishbone length. Under steady state condition( =0, =0), the balance of momentum in the wishbone around point O is given by:. =. ; (3.18) With being the force transmitted by the link determined by the balance equation for the balance bar around point C : by:.( )=. ; (3.19) The spring force is modified by the vertical displacement d of the point A, and is given = +., (3.20) With being the spring preload and its stiffness. Herein, the vertical displacement d (see Figure-20) is given by; =.., (3.21) ( ) Figure-20: Diagram of the balance bar geometry with an angular displacement for the wishbone. 14

19 CHAPTER 3- PROPOSED SOLUTIONS Finally, the actuator force is expressed by:. =..( ) +... ( ), (3.22) Force motor variation and effective stiffness of the system The motor force needed to keep the spring in place when the system is out of the ideal conditions ( =0) is represented in Figure-21. Fiure-21: Force diagram on the balance bar. In the picture, is the vertical force that balance the bar. This force is given by: = + = +. =., (3.23) The angular displacement of the wishbone results in an angular displacement in the balance bar given by: The motor force needed to keep the spring in place is given by: =., (3.24) Hence, =., (3.25) = +.. ( ).... ( ), (3.26) The effective stiffness is obtained from de derivation of actuator force equation as follow: Results from steady state analysis. = ( ) = (. )..( ), (3.27) Implementation of the equations determined in the previous section in Matlab shows the results for, and (see Figures-22 and 23). The values for the initial parameters are: =50mm, =120mm and =150mm. Figure-22: Analyses of the actuator force (right) and motor force increment (left) of Balance bar mechanism. 15

20 CHAPTER 3- PROPOSED SOLUTIONS Figure-23: Analyses for the effective stiffness of the Tetragon mechanism The results present a typical exponential characteristic. It is possible to see that this mechanism does not attend the requirements. The effective stiffness magnitude stays below 10 N/mm in a large portion of the curve and then increases to a high value (around 140 N/mm). The motor force increment assume values higher than 5000 N (reaching up to N), it is caused by the high vertical load in the joint C Realization and feasibility A CAD scheme for the balance bar is presented in Figure-24. Two solutions for the motor actuation can be seen. At first, a ball screw system transforms the rotational movement of the motor in translational motion to control the horizontal position of the pivot joint. In the second configuration, a connecting rod-crank system is responsible to control the pivot position. Figure-24: CAD sketch of the Semi-circle mechanism. The balance bar solution has one point sliding under a high pressure of contact (pivot point), and the balance bar (the lever) is working under bending moments, what could result in the lock of the pivot horizontal mobility in the mechanism. The components are not difficult to be built, but the high actuation of friction on the balance bar and its large area exposed to dirt allow the conclusion that such system is feasible, but not reliable Wishbone length variation Up to this point all presented geometries work with the same value for the wishbone length. In the device presented in this section, the length of the wishbone is controlled to vary the lever ratio. In Figure-25, is possible to see the elements that compound the wishbone. Thus, a screw connected with the electrical motor modifies the wishbone length. Figure-25:(left) Sketch of the wishbone length variation geometry. (right) a CAD view of the mechanism. 16

21 CHAPTER 3- PROPOSED SOLUTIONS Observing the Figure-8 in chapter 2, it is possible to see that this devicee follows the basic concept geometry. However, the spring position is fixed. Thus, the wishbone length l and its angular displacement are the dregrees of freedom of this system Mathematical model for the wishbone length variation The same steps used for the basic concept model can be used to write the following equations under steady state condition =0, =0. Thus, the actuator force is given by: =.( +.. ( )), (3.28) The motor force neededd to keep the spring in place is given by: =.(.+.. ( )). ( ), (3.29) Furthermore, the effective stiffness at the end of the wishbone is given by: It is important to observe that this device has the same steady state equations as the first model analyzed (the basic concept). The diagrams for these both mechanisms are the same, the only difference is that the motor force is reduced by a factor of. It is not difficult to understand why: now, the motor force has to fight against the horizontal component of the actuator force instead of the horizontal component of spring force Realization and feasibility The mechanism in discussion has some feasibility draw-backs related with the use of a screw to control the actuator. As it is required a high frequency of actuation, there is the possibility that the screw locks because of the friction. The use of ball screw is in discussion. However, the high bending moments in the wishbone make this devicee not so reliable Tetragon geometry = ( ) (. ) =., (3.30) In [10], an alternative design is presented to change the overall ratio between input and output force. The tetragon geometry, illustrated in Figure-26, works shifting the output force of the mechanism by changing the position of the arm along the arc, resulting in a controllable actuator force. Figure-26: Diagrams of the Tetragon geometry [10]. Thus, it is proposed to use one side of the tetragon as wishbone, with the output force representing the actuator force, while the spring is connected to the input force point. Hence, the wishbone angular displacement and the angular displacement of the control arm are the degrees of freedom of the system. 17

22 CHAPTER 3- PROPOSED SOLUTIONS Results from steady state analysis The steady state analysis of this device is done with the multi-body software Adams, which allows an easier measurement of effective stiffness. The results for, and are given in Figure-27, with the values for the initial parameters equal to: =50mm, h=50mm, =150mm and =150mm. Figure-27: Analyses for the tetragon mechanism: (top left) actuator force; (top right) motor force variation and (bottom) effective stiffness. The results presented still do not attend the requirements. The effective stiffness has an high amplitude of variation, and the motor force increment is higher than the basic concept. As the semi-circle mechanism, this device also seems to have advantages in terms of feasibility because it just uses link bars attached with rotational joints (which presents low influence of friction). However, a CAD scheme for this mechanism still needs to be done, to evaluate better its feasibility (to analyze, for instance, the position of the electric motor and the connection between elements). Furthermore, it geometry seems to allow a better tuning of the effective stiffness, which will be the focus of study in the next chapter Continuously variable transmission All the presented mechanisms try to create a variable suspension force. In this solution, a continuously variable transmission (CVT) is proposed to change the lever ratio of the suspension mechanism (see Figure-28(a)). An interesting CVT to use in this case would be the Nu-Vinci, detailed in reference [11]. In this mechanism the ratio is shifted by modifying the point of contact of the spheres with the input and output discs (see Figure-28(b)). Figure-28: (a) Diagram of the CVT mechanism; (b) the Nu-Vinci system. 18

23 CHAPTER 3- PROPOSED SOLUTIONS As presented in reference [11], the ratio change can be done by hand, this means, with a low force, and even if the system is not moving. That is the reason that such system has been implemented in bicycles. In Figure-29, is possible to observe a CAD scheme using such system in the suspension, with one of the arms connecting the spring while the other arm is connected to the cabin. Thus, the engine just acts to change the ratio and the motor force can be constant, as the geometry of the mechanism is not changing. However, a problem observed in all kinds of CVT s is that they use to transmit power through the friction in the contact between components. Thus, this kind of mechanism allows slip to occur. For the suspension system it would result in the lost of the spring pre-tension. This is also the most complex solution listed here and may be too expensive Conclusion Figure-29: CAD sketch of a layout using the Nu-Vinci CVT system. The requirements have not been attended so easily in this first approach. Mainly for the effective stiffness behavior, it is possible to conclude that none of the solutions present a curve within the range specified in Chapter 1. However, the Semi-circle mechanism and the Tetragon geometry seem to have advantage in terms of feasibility and reliability. Comparing the solutions, only these devices present just rotational joints and which the components do not work under bending stress. Thus, is necessary to study better how to control the behavior of the effective stiffness in these mechanisms to guaranty the requirements. Other mechanisms end up following the same principles as the geometries presented in this report, however, they were discarded for not presenting a reliable and feasible solution. 19

24 Chapter 4 Parameter analysis of the selected designs In the previous chapter, the semi-circle geometry and the tetragon mechanism are selected for further analysis. It is interesting to study how the geometric parameters of both mechanisms influence their behavior. The methodology starts with the setup of the force range required ( N). For that purpose the spring is configured with stiffness of 20 N/mm and a preload of 6400 N. The actuation force, effective stiffness and motor force are evaluated, but the main goal is to setup both devices to work within the effective stiffness range of N/mm. Both mechanisms are modeled in the multi-body software ADAMS-view with the total wishbone length of 150 mm. With the total vertical travel of 80mm at the end of the wishbone Study of the semi-circle geometry In the Semi-circle mechanism, the position of the spring is controlled by the link bar as explained in section 3.1. To observe how the geometric parameters of this device can change the effective stiffness behavior, the parameters analyzed are (see Figure-30): - The total spring length; - The vertical difference between the lower spring mount and the wishbone; - The horizontal position of the spring top mount. Figure-30: Diagram showing the parameters analyzed in Semi-circle geometry. For the total spring length, the values used are: 200, 250, 300 and 400mm (the other parameters remains unchanged). The results are presented in Figure-31, with the horizontal axes representing the spring position along the wishbone length. The increment of results in a small increment of the effective stiffness curve. However, such increment is not equal (in the right part of the curve the increment is higher).

25 CHAPTER 4- PARAMETER ANALYSIS OF THE SELECTED DESIGNS Figure-31: Comparison with different values of Lo. For the vertical difference between the lower spring mount and the wishbone the values used are: 50, 0, -50 and -100mm, and the results are presented in Figure-32. The increment of results in a decrement of the effective stiffness curve (negative values of h are under the wishbone horizontal reference line). Such decrement is higher in the left part of the curve. Furthermore, negative values for effective stiffness are obtained for positive values of h. Figure-32: Comparison with different values for h. For the horizontal position of the spring top mount the values used are: 0, 75 and 150mm, and the results are presented in Figure-33. The horizontal position of the spring top mount presents a small variation of the effective stiffness curve between 0 75mm. Just for values around 150mm the difference becomes large. However, it can be observed that this parameter can reduce the amplitude of the curve (for 75mm the curve presents the smallest amplitude). Figure-33: Comparison with different values for. 21

26 CHAPTER 4- PARAMETER ANALYSIS OF THE SELECTED DESIGNS Finally, the requirements are met using the parameters: 200 mm; 75 mm; 75 mm; For this configuration, the actuator force, motor force and effective stiffness are presented respectively in Figures-34, 35 and 36. Figure-34: Actuator force vs. the spring position along the wishbone. Figure-35: Motor force increment vs. the spring position along the wishbone. Figure-36: Effective stiffness vs. the spring position along the wishbone. It can be observed that the design requirements are accomplished using this geometry. Furthermore, the motor force is relatively low. However, the value of 200mm for (the spring length) may be too small to fit the spring. Furthermore, in the section , it is commented about the influence of the spring location in the increment of the motor force. Therefore, the semi-circle mechanism design is slightly altered to solve both problems and improve packaging as well (see Figure-37). 22

27 CHAPTER 4- PARAMETER ANALYSIS OF THE SELECTED DESIGNS Figure-37: Redesign of the semi-circle mechanism with enhanced packaging option. The spring is moved to a fixed position. The force is transfer to the spring trough a push rod system. This configuration makes the value not represent the spring length anymore but the push rod length instead Study of the tetragon geometry For the tetragon mechanism, the geometric parameters presented in Figure-38 are analyzed to observe how they can change the effective stiffness behavior of this device. The goal is to study the effect of parameters changes to evaluate if this mechanism can meet the requirements. The parameters analyzed are: - The total length of the control arm; - The vertical distance between wishbone and the input point for actuator force. - The horizontal position of the spring mount; Figure-38: Diagram showing the three parameters analyzed in the tetragon geometry and the two degrees of freedom. For the total control arm length the values used are: 100, 125, 150 and 200mm. The results are presented in Figure-39. Figure-39: Comparison with different values of Lo. 23

28 CHAPTER 4- PARAMETER ANALYSIS OF THE SELECTED DESIGNS The increment of reduces the values for effective stiffness. The major influence happens in the right part of the curve. It is possible to observe that values around 200mm can result in a negative effective stiffness. Furthermore, the curves still present high amplitude. For the vertical distance between wishbone and the input point for actuator force the values used are: 50, 0, -50 and -100mm. The results are presented in Figure-40. Figure-40: Comparison with different values of h. The increment of h increases the amplitude of the effective stiffness curve. It is possible to observe that this parameter does not show any change when the position of the control arm is -45º. Thus, the variation in the right part of the curve is higher. For the horizontal position of the spring mount the values used were: 50, 0, -50 and -100mm. The results are presented in Figure-41. Figure-41: Comparison with different values of. The variation with the parameter is almost the same than the parameter h. It makes sense due to the symmetry of the system. The angular movement of the spring makes difference between both analyzes. Finally, the requirements are met using the parameters: 120 mm; 0mm; 135 mm; Thus, the actuator force, motor force and effective stiffness are presented respectively in Figures- 42, 43 and 44. Figure-42: Actuator force vs. control arm angle for the Tetragon geometry in the final configuration. 24

29 CHAPTER 4- PARAMETER ANALYSIS OF THE SELECTED DESIGNS Figure-43: Motor force increment vs. control arm angle for the Tetragon geometry in the final configuration. Figure-44: Effective stiffness vs. control arm angle for the Tetragon geometry in the final configuration. Aparently, the requirements are attended, with the effective stiffness within the specified range. However, in Figure-42, it is possible to observe that the Fact curve for 15º intercept the curve for 0º, which means that there are 0º for which the actuator stiffness is negative. The effective stiffness is calculated around the ideal position of the wishbone ( 0, where the values for such parameter is acceptable. But for large displacements of the wishbone, the stiffness at the end of the wishbone becomes low (reaching negative for values of the control arm angle between 8 and 0 degrees). Thus, it is possible to conclude that this mechanism attends the requirements; however, the overall stiffness caracteristics is not favorable. Other values for the parameters can be used to avoid this caracteristcs, however, in that case the amplitude of the effective stiffness curve becomes too large. Furthermore, the motor force is higher if compared with the semi-circle mechanism; the reason is the length of the control arm in each device (short in the tetragon geometry), which increases the motor force magnitude. The diagram for the final tetragon design is represented in Figure-45, where is possible to observe that this device does not seem to be so packaging comparing with the semi-circle mechanism. Figure-45: (a) Tetragon geometry in its final configuration, (b) with an angular displacement. 25

30 CHAPTER 4- PARAMETER ANALYSIS OF THE SELECTED DESIGNS 4.3. Conclusion Analyzing the semi-circle and tetragon mechanisms, it is observed that the first has a good steady-state behavior. The Tetragon mechanism becomes limited to the compromise between effective stiffness and the resultant stiffness for large displacements of the wishbone. Thus, just one solution is convenient to accomplish the proposed goals. The semi-circle mechanism is feasible to manufacture, its components are simple and the use of rotational joints reduces the influence of friction (in comparison with sliders). As is possible to observe, this mechanism also presents a packaging dimensions in the configuration that accomplish the results. Herein, modifications in the semi-circle design using the push rod mechanism allowed the positioning of the spring in a fixed position, reducing the influence of its mass on the dynamic system. The next step is to analyze the possibility to fit the semi-circle geometry in a truck cabin, and evaluate the different possibilities of constructions for such system. 26

31 Chapter 5 Mechanical design The Semi-circle geometry was chosen to be implemented in the cabin suspension for its favorable characteristics. Thus, the question is how this geometry could be designed to fit in a truck cabin. The evaluation of different constructive possibilities for such system is desirable as well. This development step starts with the evaluation of the available space to fit the suspension mechanism. It is important to remember that a minimum amount of changes in the original vehicle are desired. Thus, it is deemed unsuitable to make modifications on the original layout of the truck cabin or change the position of its main components (for instance the steering system). Furthermore, in this chapter, the selection criteria for the spring and also the electric motor are specified. The configurations assembled give an overview of possibilities to use the available space, also allowing the reflection about the main constrain and problems found. Thus, the methodology tries to converge step by step for a solution that accomplishes the requirements, describing advantages and disadvantages for each one Specification of the available space The implementation in too different vehicles is investigated. For the first, the dimensions of the truck elements are obtained from measurements of a DAF truck, model CF , in the automotive laboratory of TU/e. However, some dimensions could not be provided because of impediments in reaching the front cabin suspension. For the second, a CAD model for the DAF truck XF 105 is assembled to estimate the space available in this type of heavy truck Study of the DAF CF An isometric view is represented in Figure-46, where it is possible to observe the configuration used in CF cabin suspensions ( Vx shows the forward direction of the truck). In the front, sets of air springs connect the chassis beam to the cabin. Also, a torsion bar constrains the roll movement of the cabin. In the rear, sets of air springs and damper connect the main chassis to the cabin beam as well, while a lateral spring constrains the movement to the sides. V x Figure-46: Illustration of a truck with the location of the front and rear cabin suspension.

32 CHAPTER 5- MECHANICAL DESIGN The ideal condition to analyze the space available would be a full CAD model of the front part of the truck. However, it is not possible to draw all the components, because there is no access to measure their position (several other devices are located around the front suspension like: steering, brake cylinders and some hydraulic components). However, comparing the CAD model with pictures taken from the front part of the truck, it is possible to evaluate the space occupied by these devices (see Figure-47). V x Figure-47: Picture of the front cabin spring. In the picture, is possible to observe how tight the space around the suspension spring is. Thus, the volume represented around the suspension elements on Figure-48 shows the available area to fit the proposed geometry. V x Figure-48: Space available in the front part of the truck. That volume represents the space used by the actual device plus the space occupied by the frontal beam which supports the torsion bar and the spring. The availability of space of the frontal beam is calculated considering that this component only supports the actual suspension elements (torsion bar and spring), thus, the frontal beam could also be removed together with such elements. Probably the frontal beam is important for the structural stiffness of the chassis. Furthermore, it can support other elements in the space between the two chassis beam. However, the side part of the frontal beam can be removed even though. It is not possible to advance backward or to the side, as shown in Figure-47. To the inside, the space is limited by the main chassis beam as well. However, that volume has an approximate dimension of 350 x 350 x 150 mm, which seems to be compatible with the average dimensions required by the semicircle mechanism. The possibility of using the space between the chassis beam and the cabin beam is discarded for this truck model. The reason is that some engine components are already installed there. Herein, this space is far from the point which the suspension should attach the cabin. Modification of this point is not feasible because it would change the dynamic behavior of the cabin and the initial parameters of this study as well. For the rear suspension more space is available, as can be seen in Figure

33 CHAPTER 5- MECHANICAL DESIGN Figure-49: Picture of the rear cabin suspension. The location of a hydraulic system below the suspension system does not allow the use of that space. In Figure-50, the volume around the suspension elements represents again the space available to install the Semi-circle mechanism. V x Such volume at the rear has an approximate dimension of 550 x 350 x 250 mm. The limit to the back is specified with the aim of not advance beyond the overall dimensions of the truck cabin. It is possible to conclude that the space for the front cabin suspension will constrain the size of the mechanism. In the rear, it can be interesting to follow the same layout of the front mechanism for issues of standardization Study of the DAF XF 105 Figure-50: Space available on CF model Analyzing the picture provided of the DAF truck model XF 105 [provided by dr. Igo Besselink from the TU/e], it is possible to conclude that, for this model, there is space available in the front above the chassis beam (where the actual cabin suspension is assembled). The CAD model for the XF 105 is presented in Figure-51. Figure-51: left) Picture presenting the DAF XF-105 with highlight for the cabin suspension; right) a CAD model built from the picture. 29

34 CHAPTER 5- MECHANICAL DESIGN The specifications for the space available in the model XF 105 considering just the volume occupied by the front suspension elements are 400 x 350 x 150 mm. Maybe more space is available close to the specified area, however, not knowing how other systems are positioned in the truck, it is difficult to consider the available space Actuator s elements Before describing the different solutions, it is necessary to standardize the name of each component of this mechanism (it allows a better recognition of each component functions on further commentaries). The same components can appear in different combination of position and geometry; however its function remains the same. The original concept of the Semi-circle mechanism with the name of each element is presented in Figure-52. It is shown also the main dimensions of this geometry. Figure-52: Diagram of the semi-circle mechanism with names for each component and basic dimensions. The control bar is the element attached to the motor, which sets the ratio of the mechanism. The push rod and the rocker arm are elements necessary to transfer the load to the spring, allowing the horizontal spring configuration. The wishbone is the element that attaches the chassis to the cabin beam. One additional link-bar can be necessary to allow the correct degree of freedom to the cabin depending on the positioning of the wishbone. For instance, an extra link connecting the end of the wishbone to the cabin bean is necessary if the wishbone follows the lateral direction in relation with the chassis beam (see Figure-53). Otherwise, the cabin could lose its normal degrees of freedom. Figure-53: Diagram showing the necessity of an extra link to allow the correct degree of freedom for the cabin. As presented before, the space available for the front suspension seems to follow this lateral direction; thus, probably this extra link will be present Spring specifications Using the results from the simulations presented in chapter 4, the requirements for the spring and also for the electric motor can be specified. For the spring, the specifications are given by: = 6400 N - the compression preload at neutral position ( 0); 20 N/mm - the spring stiffness; 300 mm - the spring length at the preload. 30

35 CHAPTER 5- MECHANICAL DESIGN A spring, which nearly meets the specification, is the Continental Air spring model SZ , see table-2. Table-2 Specifications of spring SZ Parameter Value Unit Load capacity at pressure = 5 bar 6200 N Natural frequency 0,97 Hz Spring stiffnes 22,7 N/mm Deflection (Height min.) 218 mm Deflection (Height max.) 378 mm Attempts to use other kinds of springs showed that the use of coil springs are not feasible to attend these specifications (would be necessary to build a large spring to accomplish, at the same time, the requirements for preload and stiffness). Also torsional springs are not feasible; their dimensions would be larger than the truck cabin. An observation about the air spring is that it can only work under compression Solutions for the front cab suspension for the CF The selection of the motor depends of how this component will be connected to the control arm. Before analyzing the choice for the motor, some solutions for the DAF truck model CF are presented. For this purpose, the motor used is a Bodine DC motor model 34B-5N with 280 W of continuous power and gear ratio of 40:1. The composition of solutions starts with the combination of motor and the wishbone. The first question that guides the design is: how the motor can be connected with the control bar? Furthermore, with the intent to reduce the inertia of the system, it is interesting to attach the motor directly with the chassis. Figure-54 presents one configuration for the motor fixed on chassis. Note that the mechanism is not complete, just the wishbone, the motor and the control bar are present (also the attachment point of the chassis bean and the extra link are presented). The intension of this approach is to show possible layouts for the mechanism, up to this step no structural analyses was made. V x Figure-54: (Solution-1A) CAD sketch showing a solution for the motor fixed on the chassis (with a 2D scheme on the left). In the solution 1A, the motor is fixed at the chassis bean, with its shaft aligned with the center of rotation of the wishbone. The transfer of power from the electric motor to the control bar is made through a connecting-rod mechanism. One problem in separating the motor from the wishbone is that the control system algorithm will have to compensate the angular displacement of the wishbone to keep the control bar at the same position. Thus, the control algorithm will become more complex. Furthermore, the connection between motor and control arm will need additional bars (in a crank connecting-rod mechanism). The possibility of connecting the control bar directly to the motor shaft depends on whether an additional reduction system would be necessary to increase the motor torque. In Figure-55, a solution attaching the motor to the wishbone is presented (again, the mechanism does not include the spring). 31

36 CHAPTER 5- MECHANICAL DESIGN V x Figure-55: (solution 2A) CAD sketch showing a solution for the motor attached directly to the wishbone. In the solution 2A, the electric motor is attached to the wishbone, and the control bar is connected directly to the motor s shaft, however, the radial load is transferred to the wishbone through bearings. The synergy between the wishbone and the motor can be observed. The main idea behind this solution is the possibility to combine the motor and the wishbone to reduce the occupied space. Thus, the wishbone would work also as the motor house. The wishbone is positioned parallel to the chassis; however, it is possible to see that the extra link is necessary to connect the wishbone to the cabin. The triangular geometry of the wishbone makes it necessary, because an off-set should be given to avoid mechanical conflict with the cabin floor. Even if the mechanism turns upside down a conflict between control bar and the cabin floor continues. Inserting the missing components (spring, push-rod, and rocker arm), and considering the space available to fit the spring, an obvious solution is to extend the mechanism along the orthogonal axis in relation to the wishbone plan. The challenge becomes to fit the push-rod elements to keep the geometric properties of the original concept. The Figure-56 presents solutions for these elements. V x Figure-56: (solution 2B) sketch with the inserting the spring, push rod, and the rocker arm to solution 2A. As presented in Figure 52 as well, the push-rod connects the control bar to the rocker arm, and then this last element transfer the movement to the spring. The original triangular format for the rocker arm is not necessary, because the spring does not work in the same plane as the wishbone. Instead, the rocker arm is transformed in a small torsion bar, also to compensate the large off-set between the spring line of actuation and the wishbone plan. Attempts of different configurations for the mechanism can be seen in Figure-57, where more suggestions for the spring and wishbone position are given. Since the extra link seems to be necessary to connect the wishbone to the cabin, the orientation of the wishbone following the lateral orientation of the truck is also an option. Furthermore, it can align the work of the spring at the same plan as the push-rod. These solutions narrowly follow the original idea (see Figure-57). The wishbone is attached to the chassis and the extra link connects the wishbone to the cabin. The main difference is that, in the figure on the right, all attachments points are at the same side; however the extra link lies far from the cabin mounting point. The difference in the location of the spring brings new challenges to fix the attachment points to the main beam of the chassis. 32

37 CHAPTER 5- MECHANICAL DESIGN V x V x Figure-57: (solutions 2C and 2D) CAD sketch of solutions that narrowly follow the previous layout designed for the Semi-circle mechanism. The solution on the left can be feasible with the introduction of a transversal beam to hold the spring. These solutions have the most compact configuration; however all these solutions do not attend the space requirements of the CF, and the main reason for it is the dimension of the spring (larger than the thickness specified). Another solution suggests changes in the rocker arm to create a different lever ratio between the spring and the push-rod, the intention is to reduce the spring thickness. Figure-58 presents one solution which follows this idea, and also includes a diagram to show the modification of the basic concept. V x Figure-58: (solution 2E) Layout using a smaller spring in a combination with a lever (replacing the rocker arm). In such configuration the lever has a ratio of 2:1 The spring selected for this application is a Firestone air spring model number 7012, its specifications are presented in table-3. Table-3 Specifications for spring Parameter Value Unit Load capacity at pressure = 5 bar 3180 N Natural frequency 1,31 Hz Spring stiffnes 13 N/mm Deflection (Height min.) 105 mm Deflection (Height max.) 295 mm However, this modification induces high bending moments on the rocker arm, reducing the reliability of this component. 33

38 CHAPTER 5- MECHANICAL DESIGN Using the air spring selected previously (SZ ), it seems difficult to fit the whole mechanism in the space defined. Thus, one possibility would be allocate the spring in other part of the truck and connect it with the push-rod by using a torsion bar. Figure-59 presents one solution that follows this idea. V x Figure-59: (solution 2F) configuration using the space available behind the tire. The torsion bar replaces the rocker arm in this solution. This solution spreads the cabin suspension system along the chassis and increase the total weight of the suspension. However, this solution attended the space constrains using an available area behind the front wheel. The only change necessary was the modification of the rocker arm in a torsion bar. Thus, it is possible to conclude that only solutions 2E and 2F attends the requirement for the front suspension of model CF. However, neither is thought to be convenient in practice Selection of the motor Since the option for positioning the electric motor is clear now, it is necessary to analyze the specifications required for the correct actuation of such component. The goal is to check if the motor selected attend the requirements and compare if there is a better option for the mechanism. For the motor selection, the boundary conditions were obtained from the analysis of force motor variation presented in chapter 4. That force motor represents the force to hold the control bar on its position. This force was measured at the end of the control bar (in the joint attached to the push rod), however, this force should be transformed in a motor torque to be used as a criteria to select the engine. Furthermore, the maximum angular speed for the control bar is necessary to be specified as well. These two parameters are used to define how much power the electric motor should have. Thus, the maximum torque required is given by: _., (5.1) With _ being the maximum motor force obtained from simulations (1500 N), and being the control bar length (200 mm). Thus, the maximum torque required is 300 N.m. Furthermore, in chapter 1, is specified the maximum value for the force motor variation (10000 N/s). Since the overall force variation of the mechanism is 6400 N, the time to complete one variation from 0 to 6400 N is given by: 0,64 s, (5.2) Herein, the total angular travel of the control bar associated with this force variation is 62 degrees. Thus, the control bar s angular speed is given by: _. / _. /0,64 1,69 rad/s, (5.3) Hence, the maximum power required by the mechanism is given by:. _ 507 watts, (5.4) There are many options of electrical motor to attend this requirement. Even the previous motor selected for the solutions with the model CF can easily attend this requirement with an increment of the 34

39 CHAPTER 5- MECHANICAL DESIGN supply current. However, the requirement for the maximum torque constrains the options. It makes necessary the use of a transmission system with a high ratio coupled to the motor. Moreover, the dimensions of the motor and the gear system influence the selection, since the space for the mechanism constrains the design. In Figure-60, some options that attend the requirements are presented beside the previous motor used (all solutions presented are combined already with a transmission system and are in the same scale). (1) (2) (3) (4) Figure-60: Comparison between options of electric motors to control the mechanism. (1): Bayside K (3150w) with Cyclo transmission (ratio 59:1); (2)Bodine 34B-5N (280w and gear ratio of 40:1); (3) Transmotec B (660w and gear ration of 70:1); (4) Transmotec PD (549w and gear ratio of 96:1). Hence, the motor selected to control the mechanism is a Parker Bayside DC motor, model K [13]. The specifications of such motor are presented on table 4. Table-4 Specifications for motor K Parameter Value Unit Maximum motor current (Imax) 75 A Maximum motor voltage (Vmax) 42 V Motor constant (Km) 0,438. / Peak torque (Tp) 14,82. Continuous torque (Tc) 4,935. Weight 2,02 kg The choice of this motor is combined with the selection of the transmission system called Cyclo, with gear ratio of 59:1. This combination has a more compact solution that attends the requirements Solution for the front cab suspension for the XF On the DAF truck model XF the position of the cabin and chassis attachment points contribute for the elimination of the extra link. Thus, one obvious position for the wishbone is replacing the anti-roll bar of the cabin suspension. Furthermore, based on the previous experience earned with the creation of solutions for the truck model CF, there is advantage in connect the motor directly to the wishbone. These observations orient the solutions for truck model XF. In Figure-61, this first design step is presented (just the wishbone and motor components are shown). Figure-61: (solution 3A) Sketch presenting the joint between cabin mount, wishbone, and chassis mount. The position of the motor on the wishbone can be seen as well. 35

40 CHAPTER 5- MECHANICAL DESIGN Some changes in the original layout of the truck cabin elements should be commented. At first, the distance between attachment points were reduced to 150 mm (the size specified for the wishbone length of the semi-circle mechanism). Moreover, a hook shape is proposed for the attachment point that connects to the chassis bean. This layout is necessary to allow the complete angular travel of the wishbone without occurring of interference between this component and the attachment arm. Note that what influences this necessity is the diameter of the motor. In Figure-62, a cross section of the mechanism is presented with the insertion of the control bar and the push rod. Figure-62: Cross section of solution 3A with the insertion of the control bars and push rod. The motor and the Cyclo transmission are positioned on one side of the wishbone while the control bar and the push rod are on the other side. This configuration allows a better use of the space available. In this solution, just one of the control bars is connected with the motor and receives power to control the mechanism. However, two control bars are used. The reason is the aim to reduce the bending moment that could act on the control bar due to the distance between the center line of the push rod and the center line of the control bar. Using two control bars it is possible to balance the system, eliminating this bend moment. The push rod consists in a bar connected to a rod end on each of its extremities. The position of the bearings allows the mechanism to transfer the load directly from the control bar to the wishbone. It was a modification created on the Cyclo transmission to reduce the occupied space. The original configuration of the Cyclo transmission is presented in Figure-63. Figure-63: Assembly example suggested by Cyclo catalog, where the out-put flange size can be compared with total size of this device. The layout suggested can also be compared with the solution proposed in Figure-62. Comparing the proposed layout of the Cyclo transmission system with the configuration used, the advantage of connecting the control bar directly to the transmission flange is clear. 36

41 CHAPTER 5- MECHANICAL DESIGN The insertion of the rocker arm and the spring complete the mechanism. In Figure-64, one solution for the XF cabin suspension is presented. Figure-64: (solution 3B) solution for the front cabin suspension of DAF truck model XF 105. The rocker arm connects the pushrod to the spring. This solution puts the spring in the horizontal position just above the chassis bean. This area is out of the space specified in section ; however, it is a feasible solution which seems to be compact. More details about the position of the truck components is required to allow a better judgment. In Figure-65, two images are presented of the mechanism in conditions of maximum angular displacement of the wishbone. Figure-65: Solution 6A with vertical displacement of the cabin. (left) bump; (right) rebound. The complete device occupies a total volume with 400 mm of height, 250 mm of length, and 300 mm of width, plus an additional space for the spring (250mm length and diameter of 170mm). Thus, this device exceeds the space specified. More details about the position of the truck components are required for a better evaluation of the XF model though Solutions for the rear cab suspension For the rear cabin suspension the main challenge is not to find space, but to find good positions to attach the elements of the mechanism. There are probably differences in the rear suspension system between the CF and XF models (related to the dimensions of the truck and the weight of the cabin); however, they have the same layout. Thus, the same solution of layout will be proposed for both models. At the rear system, particular components control the side movement of the cabin; it means that the proposed mechanism cannot constrain this degree of freedom. For this reason the extra link is present 37

42 CHAPTER 5- MECHANICAL DESIGN (it connects the wishbone to the cabin attachment point). In Figure-66, a solution for the rear system is presented. V x Figure-66: Solution for the rear cabin suspension. (left) layout of the front part of the truck showing the position of the rear suspension; (right) details of the mechanism that follows the same concept of the front suspension given for model XF. This solution for the rear fits in the available space. Furthermore, this solution could follow the same layout as in the front (just a small modification in the rocker arm is necessary to fit the spring properly) Conclusion For the CAD models built as reference, there is not so much space available in the front to fit the proposed mechanism. Also, the dimensions of the spring and motor govern the configuration of this device. Due to the motor dimensions it is concluded that the mechanism proposed is not feasible to be fitted in model CF. For the model XF, the dimensions of the proposed device exceed the dimensions occupied by the actual suspension system, however it seems to feasible for this model because of the format of the configuration reached. However, more information about the position of other systems that occupy space on the front part of the truck is necessary to better evaluate this solution. Furthermore, the selection of the electric motor is guided by the continuous torque it can reach combined with the ratio and dimensions of the correspondent gear system. Some motor attend the requirements for total power, but a large transmission system would be necessary. The combinations of criteria for torque, total power and occupied space lead to the selection of Bayside K The possibility to use the Cyclo transmission system in a different configuration than suggested in catalog also allowed a reduction on the mechanism volume. For the air spring, the diameter rules the preload capacity and the total volume defines the stiffness. Therefore, it is more difficult to find other models of different size that attends the same requirement. Finally, it is concluded that the space available on the rear allows the standardization of the solution used at front cabin suspension. Just a minimum change in the rocker arm is necessary to make it possible. 38

43 Chapter 6 Conclusion The health of truck drivers is an important issue that justifies the importance to improve the comfort for the driver. At the same time, there is a compromise between handling behavior and comfort that makes interesting the use of active suspension systems. Since the current active suspensions have a high energy consumption involved, a new concept based on a low power consumption concept is desired. The variable geometry based on Delft Active Suspension is chosen to be studied due to its low power principle. Thus, the main goal is to design a truck cabin active suspension that attends the requirements for packaging, low complexity, and within the range for the suspension parameters specified (see section 1.2.). Following the DAS principle, several mechanisms for changing lever ratio can be found on literature; however, they do not always present a feasible geometry. Most of them use translational joints, which in practice, are constructed using sliders or ball screw (solutions not reliable for application in truck cabin suspension). The results for the basic concept of DAS principle (presented in chapter 2) showed the necessity to look for mechanisms that present better behavior for the actuation force, motor force variation, and effective stiffness, with the aim to attend the requirements. The study of different geometries for possible actuators presented particular advantages for two mechanism proposed (the tetragon geometry and the semi-circle geometry both are presented in chapter 3). Such advantages are related with the 2D layout of these mechanisms, number and type of joints used, and the layout of the structural components. Comparing with the basic concept, the number of links and joints are higher; however, it is what makes these solutions feasible. Furthermore, for these mechanisms, the use of link bars allows the positioning of the spring in a fixed position, reducing the influence of its mass on the dynamic of the control system. The analysis of parameters that control the steady state behavior of the tetragon and the semicircle mechanism presented particular properties of such geometries, which allow the set-up of the geometry to accomplish the requirements for effective stiffness of the actuator. Thus, the prerequisite is reached for both mechanisms. However, for the tetragon, the effective stiffness has an undesirable behavior for large displacements of the suspension wishbone (negative values for the effective stiffness can be reached). Herein, the basic dimensions found for the semi-circle geometry is more package than the tetragon. Analysis of the space available for the DAF truck models XF and CF define the volume occupied by the actual cabin suspension system: 350 x 350 x 150 mm for the front cabin suspension of CF; 400 x 350 x 150 mm for the front cabin suspension of model XF; 550 x 350 x 250 mm for the rear cabin suspension of CF, value also used for the rear cabin suspension of the model XF). However, the solutions proposed for the front cabin suspension do not fit in the specified space. The causes are the dimensions of the spring and the motor that govern the configuration of the device. Due to these constrains it is concluded that the mechanism proposed is not feasible to be fitted in the CF model. For the XF model, the device also exceeds the space specified, however the configuration reached

44 CHAPTER 1- INTRODUCTION for this model seems to be feasible because of its compact layout. More details about the positions of the truck components are required for a better evaluation of the XF model though. The proposed device is feasible to be fitted at the rear cabin suspension of both models. Furthermore, the space available at the rear allows the standardization of the solution used at the front cabin suspension. Just a minimum changes in some components are necessary to adapt it. Comparing with the previous design presented in Chapter 1 (elpas), the volume occupied by the semi-circle mechanism is approximately 25,2 liters, while the elpas occupies approximately 27,5 liters. The main dimensions of each device are presented in Figure-67. Furthermore, the large value for the height of the elpas is not convenient to fit it in a truck cabin. Figure-67: Comparison between the dimensions of the semi-circle designed for the front suspension of the DAF truck XF model (top left), with a cross-section for the same mechanism (down left) and the elpas mechanism (top right). The specifications about the mass of the components of the semi-circle mechanism are presented on table-5. Table-5 Component s mass of the semi-circle mechanism. Component Value Unit Spring 1.500,00 g Wishbone ,00 g Control bars 1.925,00 g Push rod 850,00 g Rocker arm 1.700,00 g Motor 2.100,00 g Reduction (Cyclo) 2.600,00 g TOTAL ,00 g The mechanism attends the requirements specified. From table-5, it is possible to conclude that the wishbone is the component that influences more the mass of this system (representing 52% of the system). The combination of electric motor with the Cyclo reduction represents 21% of the system s weight. 40