TRANSILVANIA UNIVERSITY OF BRASOV MECHANICAL ENGINEERING FACULTY AUTOMOTIVE AND TRANSPORTATION DEPARTMENT AUTOMOTIVE ENGINEERING GRADUATE PROJECT… [303375]

TRANSILVANIA UNIVERSITY OF BRASOV

MECHANICAL ENGINEERING FACULTY

AUTOMOTIVE AND TRANSPORTATION DEPARTMENT

AUTOMOTIVE ENGINEERING

GRADUATE PROJECT

Graduate:

Bogdan Ioan ZAMFIR

Project coordinators:

Prof. Dr. Ion PREDA

Dr. Gabriel DIMA

BRASOV

2017

[anonimizat]:

Bogdan Ioan ZAMFIR

Project coordinators:

Prof. Dr. Ion PREDA

Dr. Gabriel DIMA

IN COLLABORATION WITH SCHAEFFLER ROMANIA

2017

I want to thank to Schaeffler Romania for the given opportunity to develop the graduate project in the company.

For their unconditional support, I want to thank to Mr. [anonimizat]. [anonimizat]. Gabriel Dima.

[anonimizat] a lot of interesting and important subjects have been treated.

The current thesis comprises two parts in which there have been dealt two separate subjects.

[anonimizat]-[anonimizat], [anonimizat], [anonimizat]. Furthermore, [anonimizat].

[anonimizat] a new design for a [anonimizat]. The main steps that were followed in order to design the coupling were: [anonimizat], strength calculus (realized in parallel with the 3D modelling), [anonimizat].

Overall, [anonimizat], as well as a new innovative technical solution regarding safety couplings in industrial engineering.

[anonimizat]’s suspension

The suspension of modern vehicles need to satisfy a number of requirements whose aims partly conflict because of different operating conditions (loaded/ unloaded, acceleration/braking, level/[anonimizat]/cornering).

The forces and moments that operate in the wheel contact area must be directed into the body. [anonimizat] a [anonimizat], the angle of the steering axis.

[anonimizat] (kinematic wheel) is required for reasons of ride comfort. [anonimizat] a [anonimizat], for which the most rigid wheel suspension is required. This requirement is undermined as a result of the necessary flexibility that results from disturbing wheel movements generated by longitudinal forces arising from driving and braking operations.

For the purpose of ensuring the optimum handling characteristics of the vehicle in a steady and transient state, the wheels must be in a defined position with respect to the road surface in order to generate the necessary lateral forces. The size of the lateral forces are determined by specific toe-in and camber changes of the wheels depending on the jounce and movement of the body as a result of the axle kinematics (roll steer) and operative forces (compliance steer). This make it possible for specific operation conditions such as load and traction to be taken into consideration. By establishing the relevant geometry and kinematics of the axle, it is also possible to prevent the undesirable diving or lifting of the body during braking or accelerating and to ensure that the vehicle does not exhibit any tendency to oversteer and displays predictable transition behavior for driver [1].

Other requirements:

Independent movement of each of the wheels on an axle (not guaranteed in the case of rigid axles);

Small, unsprung masses of the suspension in order to keep the wheel load fluctuation as low as possible;

The introduction of the wheel forces into the body in a manner favorable to the flow forces;

The necessary room and expenditure for construction purposes;

Behavior with regard to the passive safety of passengers and other road users;

Costs.

Double wishbone suspension

The double wishbone suspension consists of two transverse links (control arms) either side of the vehicle, which are mounted to rotate on the frame, suspension subframe or body and, in the case of the front axle, are connected on the outside to the steering knuckle or swivel heads via ball joints. The greater the effective distance c between the trans- verse links (Fig.1), the smaller the forces in the suspension control arms and their mountings become, i.e. component deformation is smaller and wheel control more precise [1], [23].

Fig. 1 Double wishbone suspension configuration [1]

On front independent wheel suspensions, the lateral cornering force FY,W,f causes the reaction forces FY,E and FY,G in the links joining the axle with the body. Moments are generated on both the outside and the inside of the bend and these adversely affect the roll pitch of the body. The effective distance c between points E and G on a double wishbone suspension should be as large as possible to achieve small forces in the body and link bearings and to limit the deformation of the rubber elements fitted.

The main advantages of the double wishbone suspension are its kinematic possibilities. The positions of the suspension control arms relative to one another can determine both the height of the body roll center and the pitch pole. Moreover, the different wishbone lengths can influence the angle movements of the compressing and rebounding wheels, i.e. the change of camber and, irrespective of this, to a certain extent also the track width change. With shorter upper suspension control arms the compressing wheels go into negative camber and the rebounding wheels into positive. This counteracts the change of camber caused by the roll pitch of the body. The vehicle pitch pole O is located behind the wheels on the front axle and in front of the wheels on the rear axle. If Or can be located over the wheel center, it produces not only a better anti-dive mechanism, but also reduces the squat on the driven rear axles (or lift on the front axles). These are also the reasons why the double wishbone suspension is used as the rear axle on more and more passenger cars, irrespective of the type of drive, and why it is progressively replacing the semi-trailing link axle [1].

Fig. 1.1. Volkswagen light commercial vehicle front axle [1]

MacPherson strut suspension

The MacPherson strut is a further development of double wishbone suspension. The upper transverse link is replaced by a pivot point on the wheel house panel, which takes the end of the piston rod and the coil spring. Forces from all directions are concentrated at this point and these cause bending stress in the piston rod. To avoid detrimental elastic camber and caster changes, the normal rod diameter of 11 mm (in the shock absorber) must be increased to at least 18 mm. With a piston diameter of usually 30 mm or 32 mm the damper works on the twin-tube system and can be non-pressurized or pressurized.

Fig. 1.2. McPherson strut configuration [1]

In Fig.1.2. is presented the rear view of the left-hand side of the McPherson front axle on the Opel Omega (1999) with negative kingpin offset at ground (scrub radius) rs and pendulum-linked anti-roll bar. The coil spring is offset from the MacPherson strut to decrease friction between piston rod 2 and the rod guide. Part 2 and the upper spring seat 9 are fixed to the inner wheel house panel via the decoupled strut mount 10. The additional elastomer spring 11 is joined to seat 9 from the inside, and on the underside it carries the dust boot 12, which contacts the spring seat 3 and protects the chrome-plated piston rod 2. When the wheel bottoms out, the elastomer spring rests on the cap of the supporting tube 1. Brackets 4 and 13 are welded to part 1, on which the upper ball joint of the anti-roll bar rod 5 is fastened from inside. Bracket 13 takes the steering knuckle in between the U-shaped side arms.

The upper hole of bracket 13 has been designed as an elongated hole so that the camber can be set precisely at the factory.

A second-generation double-row angular (contact) ball bearing (item 14) controls the wheel.

The ball pivot of the guiding joint G is joined to the steering knuckle by means of clamping forces. The transverse screw 15 grips into a ring groove of the joint bolt and prevents it from slipping out in the event of the screw loosening.

The subframe 6 is fixed to the body. In addition to the transverse control arms, it also takes the engine mounts 8 and the back of the anti-roll bar 7. The drop center rim is asymmetrical to allow negative wheel offset (not shown) at ground (scrub radius) [1].

Advantages:

Long spring travel;

Three bearing positions no longer needed;

Better design options on the front crumple zone;

Space at the side permitting a wide engine compartment;

Makes it easy to fit transverse engines.

Disadvantages:

Less favorable kinematic characteristics;

Introduction of forces and vibrations into the inner wheel house panel and therefore into a relatively elastic area of the front end of the vehicle;

It is more difficult to insulate against road noise – an upper strut mount is necessary, which should be as decoupled as possible ;

The friction between piston rod and guide impairs the springing effect;

In the case of high-mounted rack and pinion steering, long tie rods and, consequently, more expensive steering systems are required ;

Greater sensitivity of the front axle to tyre imbalance and radial ;

Greater clearance height requirement;

Sometimes the space between the tyres and the damping element is very limited.

Multi-link suspensions

A form of multi-link suspension was first developed by Mercedes-Benz in 1982 for the 190 series. Driven and non-driven multi-link front and rear suspensions have since been used.

Up to five links are used to control wheel forces and torque depending on the geometry, kinematics, elasto- kinematics and force application of the axle. As the arrangement of links is almost a matter of choice depending on the amount of available space, there is extraordinarily a wide scope for design. In addition to the known benefits of independent wheel suspensions, with the relevant configuration the front and rear systems also offer the following advantages [4], [21]:

Free and independent establishment of the kingpin offset, disturbing force and torque developed by the radial load;

Considerable opportunities for balancing the pitching movements of vehicles during braking and acceleration (up to more than 100% anti-dive, anti- lift and anti-squat possible);

Wide scope for design with regard to elastokinematic compensation from the point of view of (a) specific elastokinematic toe-in changes under lateral and longitudinal forces and (b) longitudinal elasticity with a view to riding comfort (high running wheel comfort) with accurate wheel control.

Disadvantages:

Increased expenditure as a result of the high number of links and bearings;

Higher production and assembly costs;

The possibility of kinematic overcorrection of the axle resulting in necessary deformation of the bearings during vertical or longitudinal movements;

Greater sensitivity to wear of the link bearings;

High requirements with regard to the observation of tolerances relating to geometry and rigidity.

Fig. 1.3. BMW 5 series (E39, 1996) rear multi-link suspension [1]

In Fig.1.3 is presented the multi-link rear suspension of the BMW 5 series (E39, 1996). For the first time in large-scale car production, mainly aluminum is used for the suspension system derived from the geometry of the BMW 7 series.

The subframe (rear-axle support) (1), produced from welded aluminium tubes, is attached to the bodywork by means of four large rubber mounts (2). These are soft in a longitudinal direction for the purposes of riding comfort and noise insulation and rigid in a transverse direction to achieve accurate wheel control. The differential gear also has compliant mounts (3). The wheel carrier is mounted on a U- shaped arm (5) at the bottom and on the transverse link (7) and inclined guide link (8) at the top. As a result of this inclined position, an instantaneous centre is produced between the transverse link and guide link outside the vehicle which leads to the desired brake understeer during cornering and the elastokinematic compensation of deformation of the rubber bearings and components. The driving and braking torque of the wheel carrier (11) is borne by the ‘integral’ link (9) on the swinging arm (5), which is subject to additional torsional stress as a result. This design makes it possible to ensure longitudinally elastic control of the swinging arm on the guide bearing (10) for reasons of comfort, without braking or driving torque twisting the guide bearings as would be the case with torque borne by pairs of longitudinal links. The stabilizer behind presses on the swinging arm (5) by means of the stabilizer link (6), whereas the twin-tube gas-pressure shock absorber, whose outer tube is also made of aluminium, and the suspension springs provide a favourably large spring base attached directly to the wheel carrier (11). For reasons of weight, the wheel discs are also made of aluminium plate. The wheel carrier is made of shell cast aluminium. The rear axle of the station wagon BMW Tourer is largely similar in design. However, the shock absorber extends from the U-shaped swinging arm in order to allow for a wide and low loading area [1].

Springs

The spring is the main component of the suspension system, and four types are primarily in use today: leaf springs, torsion bars, coil springs, and pneumatic (air) springs.

Leaf springs

Most early cars used this type of spring because leaf springs were used extensively on horse-drawn carriages, and early designers had some experience with them. The leaf spring shown in the figure below is a multi-leaf type. This type of spring is made of a single elliptical spring with several smaller leaves attached to it with clamps.

The leaves also are fixed rigidly by the center bolt, which prevents individual leaves from moving off-center during deflection. The additional leaves make the spring stiffer, allowing it to support greater loads. Furthermore, as the spring deflects, friction is generated between the leaves, resulting in some damping capability. Leaf springs also provide fore-and-aft location, as well as some lateral location, for the axle. Although leaf springs are simple and cheap, they tend to be heavy. Leaf springs also weaken with age and are susceptible to sag.

Fig. 2 Multi leaf spring [2]

Torsion bars

The torsion bar is a circular steel rod made of spring steel. One end of the rod is anchored to the frame, and loading is pure shear due to torsion. The figure below shows an example of a torsion bar. The torsion bar has very little inherent damping and therefore must be used in conjunction with dampers. As long as the bar remains in the elastic region, torque resistance will return the bar to its normal position upon unloading. The primary disadvantage of torsion bars is the axial space required for installation [2], [25].

Fig. 2.1. Torsion bar [25]

Fig. 2.2. Torsion bar suspension [2]

Coil Springs

Coil springs are basically torsion bars that have been wrapped into a coil. The figure below shows an example of a coil spring suspension. Similar to torsion bars, coil springs have little to no inherent damping and require the use of dampers. Coil springs are used widely in automotive applications due to their compact size. However, coil springs are not capable of providing any location of the axle; thus, they require control arms to limit longitudinal and lateral suspension motion.

Fig. 2.3. Coil Spring Suspension [2]

Coil springs parameters

Mean coil diameter, D: The center-to-center distance of the wire across the coil.

Wire diameter, d.

Pitch, p: The distance between successive coils on an uncompressed (free) spring.

Spring index, C: C = D/d and normally is greater than 3.

Spring rate, k: k = F/δ, where F is applied load and δ is deflection.

Active coils: The number of coils not touching the support [27].

Fig. 2.4. Coil spring dimensions [26]

Materials and technology used in springs manufacturing

Cold winding

Wire up to 0.75 in (18 mm) in diameter can be coiled at room temperature using one of two basic techniques. One consists of winding the wire around a shaft called an arbor or mandrel. This may be done on a dedicated spring-winding machine, a lathe, an electric hand drill with the mandrel secured in the chuck, or a winding machine operated by hand cranking. A guiding mechanism, such as the lead screw on a lathe, must be used to align the wire into the desired pitch (distance between successive coils) as it wraps around the mandrel.

Alternatively, the wire may be coiled without a mandrel. This is generally done with a central navigation computer (CNC) machine.

The wire is pushed forward over a support block toward a grooved head that deflects the wire, forcing it to bend. The head and support block can be moved relative to each other in as many as five directions to control the diameter and pitch of the spring that is being formed.

For extension or torsion springs, the ends are bent into the desired loops, hooks, or straight sections after the coiling operation is completed [26].

Fig. 2.4.1. Cold winding process [16]

Hot winding

Thicker wire or bar stock can be coiled into springs if the metal is heated to make it flexible. Standard industrial coiling machines can handle steel bar up to 3 in (75 mm) in diameter, and custom springs have reportedly been made from bars as much as 6 in (150 mm) thick. The steel is coiled around a mandrel while red hot. Then it is immediately removed from the coiling machine and plunged into oil to cool it quickly and harden it.

Fig. 2.4.2. Hot winding process [17]

Hardening

Whether the steel has been coiled hot or cold, the process has created stress within the material. To relieve this stress and allow the steel to maintain its characteristic resilience, the spring must be tempered by heat treating it. The spring is heated in an oven, held at the appropriate temperature for a predetermined time, and then allowed to cool slowly. For example, a spring made of music wire is heated to (260°C).

Grinding

If the design calls for flat ends on the spring, the ends are ground at this stage of the manufacturing process. The spring is mounted in a jig to ensure the correct orientation during grinding, and it is held against a rotating abrasive wheel until the desired degree of flatness is obtained. When highly automated equipment is used, the spring is held in a sleeve while both ends are ground simultaneously, first by coarse wheels and then by finer wheels. An appropriate fluid (water or an oil-based substance) may be used to cool the spring, lubricate the grinding wheel, and carry away particles during the grinding [25].

Fig. 2.4.4. Grinding process [18]

Shoot peening

This process strengthens the steel to resist metal fatigue and cracking during its lifetime of repeated flexings. The entire surface of the spring is exposed to a barrage of tiny steel balls that hammer it smooth and compress the steel that lies just below the surface.

Fig. 2.4.5. Shoot peening process [19]

Setting

To permanently fix the desired length and pitch of the spring, it is fully compressed so that all the coils touch each other. Some manufacturers repeat this process several times.

Fig. 2.4.6. Setting process [20]

Coating

To prevent corrosion, the entire surface of the spring is protected by painting it, dipping it in liquid rubber, or plating it with another metal such as zinc or chromium. One process, called mechanical plating, involves tumbling the spring in a container with metallic powder, water, accelerant chemicals, and tiny glass beads that pound the metallic powder onto the spring surface.

Alternatively, in electroplating, the spring is immersed in an electrically conductive liquid that will corrode the plating metal but not the spring. A negative electrical charge is applied to the spring. Also immersed in the liquid is a supply of the plating metal, and it is given a positive electrical charge. As the plating metal dissolves in the liquid, it releases positively charged molecules that are attracted to the negatively charged spring, where they bond chemically. Electroplating makes carbon steel springs brittle, so shortly after plating (less than four hours) they must be baked at 325-375°F (160-190°C) for four hours to counteract the embrittlement [26].

Fig. 2.4.7. Coating process [21]

Spring materials

Springs are usually made from alloys of steel. The most common spring steels are music wire, oil tempered wire, chrome silicon, chrome vanadium, and 302 and 17-7 stainless. Other materials can also be formed into springs, depending on the characteristics needed. Some of the more common of these exotic metals include beryllium copper, phosphor bronze, Inconel, Monel, and titanium. Titanium is the strongest material, but it is very expensive. Next come chrome vanadium and chrome silicon, then music wire, and then oil tempered wire. The stainless and exotic materials are all weaker than the rest.

The spring materials are given below:

Hard-drawn wire: This is cold drawn, cheapest spring steel. Normally used for low stress and static load. The material is not suitable at subzero temperatures or at temperatures above 1200C.

Oil-tempered wire: It is a cold drawn, quenched, tempered, and general purpose spring steel. It is not suitable for fatigue or sudden loads, at subzero temperatures and at temperatures above 1800C.

Chrome Vanadium: This alloy spring steel is used for high stress conditions and at high temperature up to 2200C. It is good for fatigue resistance and long endurance for shock and impact loads.

Chrome Silicon: This material can be used for highly stressed springs. It offers excellent service for long life, shock loading and for temperature up to 2500C.

Stainless steel: Widely used alloy spring materials.

Phosphor Bronze / Spring Brass: It has good corrosion resistance and electrical conductivity. It is commonly used for contacts in electrical switches. Spring brass can be used at subzero temperatures.

Analysis of the dynamic performances of an off-road vehicle

In this section it has been used the knowledge from vehicle’s dynamics in order to determine the dynamic performances of a four wheel drive vehicle.

This analysis consists in the following:

Realizing the vehicle’s critical dimensions drawing.

Realizing the kinematic scheme of the vehicle.

Computation of the vehicle’s center of gravity position (loaded and unloaded state).

Grade ability – the computation of the ramp angle the vehicle can climb.

Traction characteristics – the computation of all important characteristics an engine features (power, torque, engine speed, hourly fuel consumption, etc.).

Braking performances – analyzing how the vehicle is behaving in the case of an emergency braking function of the quality of the road surface, driver reaction time, braking system reaction time, speed, etc.

Ackermann steering

Stability – cornering ability of the vehicle – the computation of the skidding critical speed and rollover critical speed function of the grip coefficient.

Ride behavior of the vehicle – analyzing the oscillations resulted after the vehicle passes over a bump or a pothole in both states, loaded and unloaded.

All these drawings and calculations have been made using Autodesk AutoCAD Mechanical and PTC Mathcad Prime 3.1.

All the calculations were made using information from [3], [4], [5], [6], [7], [15], [22].

Vehicle’s critical dimensions

Kinematic scheme

Computation of the vehicle’s center of gravity

Grade ability

Traction ability

Braking ability

Ackermann steering

Ackermann’s conditions for steering:

Vehicle’s dimensions:

w=2500 mm – wheelbase

T=1572 mm – tread

a=1300 mm – the distance between the front axle and the center of gravity

c1=875 mm – front console

Tyre dimensions:

215/55 R17

Bpn=215 mm

Asppn=0.55

Dj=17 in

The maximum steering angle of the inner wheel:

The maximum steering angle of the outer wheel:

The outer turning radius corresponding to the rear axle:

The inner turning radius corresponding to the rear axle:

The mean turning radius corresponding to the rear axle:

The outer turning radius corresponding to the front axle:

The maximum steering angle corresponding to the vehicle’s center of gravity:

The turning radius – corresponding to the center of gravity:

The minimum (interior) turning radius:

The maximum (exterior) turning radius:

The track width:

The off-tracking:

Cornering ability

Ride behavior

Kinematics of the suspension system

The vehicle on which the kinematics of the suspension system has been made, features a MacPherson-type suspension system.

The analysis has been made in CATIA V5. Using the results obtained from the ride behavior section, these were simulated in AMESIM.

It was considered the displacement of the vehicle in three scenarios:

Travelling on flat surface (highway)

Fig. 3.1. Flat surface

Travelling over a bump

Fig. 4.2. Bump

Travelling over a pothole

Fig. 4.3. Pothole

Further, it has been performed a simulation using the quarter-car model, from which resulted the displacement of the body and the displacement of the wheel.

Fig. 4.4. The quarter-car model representation in AMESIM

Results:

Fig. 4.4. Body displacement

SAFETY INERTIAL MOBILE COUPLING

General information about couplings

The couplings establish a permanent or intermittent link between two consecutive elements of a kinematic chain, in order to transmit torque and rotation.

Besides their main function of transmitting the torque and the rotation, the couplings can perform other functions too [8]:

compensating the deviations of the elements linked by the coupling, which are due to errors of manufacturing and assembling;

connecting some shafts with parallel or concurrent axes;

protecting the transmission from shocks and vibrations ;

limiting the transmitted load;

limiting the rotations;

disconnecting the transmission when the direction of rotation is changed;

commanded disengaging of the link between the elements.

Classification

There are several criteria by which couplings can be classified:

By the physical process of transmitting the load:

Mechanical couplings (direct contact or friction);

Hydraulic couplings;

Electric couplings.

By the continuity of transmitting the load:

Permanent couplings;

Intermittent couplings.

By the possibility of compensating the deviations when assembling the coupling or during its functioning:

Fixed couplings;

Mobile couplings;

By the shape of friction surfaces:

Plane;

Conical;

Cylindrical.

One of the most used classifications, based on some of the criteria presented above, is shown in fig. 6.1:

Fig. 6.1 The classification of couplings [8]

Load calculus

During functioning, a complex system of loads acts on the coupling. Except to the torque, that has to be transmitted, on the coupling also act:

internal loads, caused by the inertia of the elements;

shocks or vibrations, caused by the working conditions of the transmission;

overloads, caused by the deformation of the elements; usually as a result of the deviations of the shafts;

supplementary loads caused by the friction between the mobile elements of the coupling [13].

In fig. 7 is shown the general variation of the transmitted torque, during functioning. The characteristic points and stages of the transmitted torque variation are:

a – the starting shock;

b – passing in the stationary regime;

c – the stationary regime;

d – the stopping shock.

Fig. 7 The variation of the transmitted torque [14]

In the couplings design the torque used for dimensioning is called calculus torque and is established with the relation:

[14]

where:

Ks – safety coefficient of overload (>1);

Mtn – nominal transmitted torque, calculated with the following relation:

, where:

P[kW] – the power of the motor;

N[rpm] – the functioning rotation of the coupling.

The coefficient Ks is determined using the experimental data and is usually obtained as a product of factors for which the values are shown in standards or in the manufacturer’s catalogues. The factors are taking into consideration the type of the driving machine, the type of the driven machine (the working conditions) and the functioning regime of the transmission.

The couplings can be chosen from standards or from the catalogues of specialized manufacturers, depending on the given torque Mt cat, following the condition Mt cat Mtc [14].

Intermittent couplings

Intermittent couplings are characterized by the fact that they offer the possibility of detaching the link element of the shafts, starting from an exterior command or automatically [8].

The main requirements of intermittent couplings are:

Capacity of transmitting the torque;

Safe coupling and decoupling, at command or automatically;

Minimum dimensions, weight and moment of inertia;

High reliability;

Easy maintenance;

Simple construction;

Low cost.

Safety Couplings

The safety couplings are represented by the automated intermittent couplings which are able to limit the transmitted load. Also, they have the function of limiting the transmitted torques in case of the appearance of overloads during the working process. Thus, the overloads of the elements from the kinematic chain and their damage are avoided [8].

The dynamic overloads can appear occasionally or periodically, especially at fast mechanisms, during accelerations or decelerations of large inertial masses or when a mechanism gets blocked. At slow mechanisms, the danger of overloads is reduced due to the small kinetic energy of the system.

The static overloads appear due to a high loading of the driven machine, both at fast mechanisms and slow ones.

The overloads can lead to the damage of the machine or to its complete destruction [13].

Features of the safety couplings:

Reliability;

Safe functioning;

Precision of torque limitation at a certain imposed value;

Ease at uncoupling;

Possibility of torque adjustment;

Capacity of automated restoration of the kinematic chain, after the overload stops.

According to the main constructive and functional criteria, the classification of the safety couplings is presented below:

Fig. 8.1 The classification of safety couplings [8]

Working regimes of the safety couplings

In order to fulfill both their main role, of transmitting the torque, and the specific one of limiting the torque, three distinct functional situations can be met during the functioning of safety couplings. These situations are defined by the value of the nominal torque that must be transmitted by the coupling and by the value of the overloads [13].

The completely coupled regime corresponds to the situation in which the transmitted torque Mt tr, is smaller than the maximum torque Mt 0, possible to be transmitted by the coupling, in this working situation: Mt = Mt tr Mt 0.

The uncoupling process corresponds to the situation in which, as the transmitted torque Mt tr overrides the maximum torque Mt 0, a relative rotation motion between the half couplings appears, beginning this process. The variation of the torque Mt = Mtd (Mt decoupling) in this process is depending on the type of the safety coupling that equips the transmission.

For safety couplings with interruption of transmitting the torque (couplings with breaking pins) the variation of the torque Mtd is shown below:

Fig. 8.1.1 Torque variation for safety couplings with the interruption of the torque [8]

The variation of the transmitted torque Mt tr , during reaction at overload, will have different shapes, depending on the dynamic characteristics of the transmission and of the coupling. At the value Mtd=Mt 0, the uncoupling of the two half couplings is produced almost instantaneously and from the torque Mt tr =Mt 0, the torque at the driven coupling reaches the value Mtd=Mtr= 0

For safety couplings with intermittent transmission of the torque (couplings with claw dogs, with balls etc.), the variation of the torque Mtd is presented below:

Fig. 8.1.2 Torque variation for safety couplings with intermittent transmission of the torque [8]

From the moment in which Mt tr=Mt 0 until reaching the maximum value of the decoupling torque Mtd, there is an area of instability, where incomplete processes of coupling and decoupling could appear. After this area is passed, the torque Mtd drops at the value Mtr, which is maintained until a new coupling process.

The couplings with continuous transmission of the torque (the transmission of the torque is performed through friction) have the optimum shape of the variation of the decoupling torque Mtd, shown in the figure below:

Fig. 8.1.3 Torque variation for safety couplings with continuous transmission of the torque [8]

After passing the area defined by the reaction time at overload, the torque Mtd decreases, continuously from Mt 0 to Mtr , where it gets stabilised. The shape of the variation curve of the torsion moment transmitted by the coupling, until Mt tr=Mt 0, depends on the constructive characteristics of the coupling (the compressing system, the friction materials etc.). The relatively high value of the remanent torque Mtr explains the wide utilisation of these couplings in the construction of machines.

The laws of variation Mt = Mt (t), shown for the three types of torque transmission are theoretical. The torque Mt tr does not have a constant value. Due to the dynamic characteristics of both driving and driven machines, there is a field of variation for the transmitted torque, which is hatched in figures 8.1.1, 8.1.2, 8.1.3 [8].

The coupling process is characterised by the equalization of the angular speeds of the two half couplings due to the decrease of torque from the transmission. This process is featured only for the safety couplings with intermittent or continuous transmission of the torque. In both cases at the end of the coupling process, the completely coupled working situation is obtained.

In case of safety couplings with interruption of transmitting the torque (breaking pins), restoring the link from the shafts is obtained by replacing the broken pin.

The rational utilisation of the driven machine imposes to the transmission the insurance of a minimum level of the transmitted load by the safety coupling, without starting the uncoupling process.

This level is established depending on the overloads that appear at the driven mechanism of the driven machine and which becomes the calculus torque Mtc for the safety coupling.

In the case when the safety couplings are designed in such a way that the maximum transmitted torque, without starting the uncoupling process, to be equal with Mtc, the working of the coupling in the area of the values near to Mtc becomes unstable, with frequent decouplings and couplings. As a result, the minimum level of the load in transmission that would allow the rational utilisation of the driven machine is not permanently ensured. This explains the design of the safety couplings with limit torque, determined with relation:

[8]

JAKOB safety couplings

JAKOB safety couplings are designed as nominal break points or as overload protection in a direct or indirect drive train.

The heart of the safety coupling is a highly precise, sturdy disengagement mechanism with steel balls as spring loaded positive locking elements.

The drive torque is guided into the centrally arranged hub via a frictional, backlash-free radial clamping hub or conical bush connection. The hub is designed as a ball cage and serves for fitting the flange ring and the shift disk. Special cup springs press the balls over the shift disk into hardened countersunk holes of the flange ring.

In normal operation, the drive torque is transferred without backlash into the flange ring. For further transfer of the torque and speed, a choice of compensation elements (metal bellows, elastomer spider), a gear or pulley or an appropriate connection flange is fit to the flange ring.

If the set disengagement torque is exceeded in the case of crash or collision, the flange ring turns in relation to the cage hub and the balls are abruptly pushed out of the holes. The drive train is cut-off within a few milliseconds [23].

Fig. 8.2 JAKOB safety coupling [23]

The backlash-free ball locking mechanism

The preload of the hardened and polished steel balls between the ball cage, the hub, and the detents of the flange ring ensures a backlash-free torque and angular motion transfer with high torsional stiffness. The mechanism is effective in reverse direction as well (for clockwise or counter-clockwise operation).

The digressive spring characteristic

The function of the safety coupling is influenced substantially by the cup springs, developed specifically for this application. Due to its operation in the degressive range, the spring force drops with increasing spring travel (shifting path), whereby the torque drops immediately on response. With conventional spring loaded torque limiters on the other hand, springs are stressed even further and the spring force as well the disengagement torque increase considerably before the actual disengagement takes place, leading to additional damage. This results in an undefined function between response and disengagement [23].

Requirements and working principles

In this section have been presented the main requirements imposed for the coupling, as well as some of its main features and working principles.

Requirements

To withstand at a transmitting torque of 1000 Nm;

The design – as compact as possible;

Bidirectional taking over the loads;

Fields of application: industrial engineering;

To protect transmissions and other important machines’ components in the case of an overload.

Working principle

The coupling works on the base of the inertia principle. As shown in figure 9, for high speed the ball overlaps the channel, while for low speeds, the ball falls into the channel.

Fig. 9 The inertia principle illustration

The transition from the coupled state to the uncoupled state is done by a hook which, depending on the speed of the internal shaft, will overlap the channel or will be attached to it, as shown in figure 9.1.:

Fig. 9.1. Low speed vs. high speed scenario

The threshold coupling time is 50 milliseconds. That means if the internal shaft is rotating with a very high speed that would require a coupling process of less than 50 milliseconds, the hook will not engage (slip occurs) with the internal shaft and, this way, the power flow will not be transferred farther. If the coupling process is achieved in more than 50 milliseconds, the hook will engage with the internal shaft and the power flow will be continuous.

In figure 9.2., it is presented the core of the coupling and its main components:

In the housing (1) there are fixed two sets of pin01 (2) on which there are attached the hooks (3). In the same way, the internal shaft (4) features two sets of pins (5), rolls (6) and blocking studs (7). The roll (6) rotates freely on the pin (5) which is fixed in the internal shaft.

There have been used two sets of hooks in order to take the loads in both directions: clockwise and counterclockwise.

The profile of the hook (3) has been designed in such a way that provides proper disengagement. Furthermore, the free rotation of the roll (6), which acts a support element, enables the disengaging process.

In the figure 9.3.there are presented the other important components of the coupling and the power flow.

The hook (3) is always pressed on the internal shaft via a torsion spring (8) which is fixed between the hook and the cap (9).

The housing (1) is rotating freely with respect to the internal shaft (4) by using needle roller bearings (10). These do not need any axial locating elements due to the presence of the toroidal springs (11) which are exerting a constant high pressure between the housing and the internal shaft.

The power flow follows the path: internal shaft – blocking stud – hook – housing – head cap screws – cap.

The coupling has been designed in PTC Creo Parametric 2.0 and in Autodesk AutoCAD Mechanical, using information from [10], [11].

Kinematic scheme

The kinematic scheme of the coupling illustrates 3 scenarios:

Engaged;

Disengaging process;

Disengaged.

Fig.10 Continuous power flow Fig. 10.1. Disengaging process

Fig.10.2. Interrupted power flow

Strength calculus

In this section is presented the strength calculus for each part subjected to stresses.

The coupling’s design has been realized in parallel with this calculation so that, in the long run, it has been developed a very robust and durable assembly.

The calculation has been performed using information from [8], [9], [10], [11], [12].

All the calculations have been made using PTC Mathcad Prime 3.1.

Splined shaft

Internal shaft

Housing

Blocking stud

Pin01

Hook

Hexagon socket head cap screws

Dynamic calculus

This calculation aimed the determination of the elastic moment acting on the hook, as well as the main dimensions of the torsion spring.

CAD design

In this section is presented the 3D design of the coupling and the main important sections.

Fig. 13 Isometric view

Fig. 13.1. Section A-A

Fig. 13.2. Section B-B

Fig. 13.3. Section C-C

Applications

The designed coupling can be used as a safety element in the following cases:

Machines transmissions where shocks can occur or where big inertial masses exist;

Machines transmissions that process inhomogeneous materials (implements);

Automatic machines transmissions, where the permanent control of the functioning is impossible;

Kinematic chains, where the protection of the transmission from the electric motor is impossible;

All transmissions where the cost of the over dimensioning is higher than the cost of a good safety coupling.

Conclusions

During the analysis of the dynamic performances of the off-road vehicle, the following facts were noticed:

The vehicle can climb the highest ramp compared with front wheel drive and rear wheel drive vehicles.

The breaking performance, as well as the cornering ability, highly depends on the road surface.

The vehicle features a neutral steering behavior compared with front wheel drive vehicles which feature understeering behavior and rear wheel drive vehicles which feature oversteering behavior.

The passengers’ comfort is highly dependent on the suspension system, by the way it softens the ride when travelling in different road conditions.

The design of the safety inertial mobile coupling has been a challenging project in which it was essential to work in the same time with the design and the strength calculus. This finally led to an innovative, feasible and robust solution for safety couplings

Nowadays, the inertial couplings have very low applications in industrial engineering. By developing this type of coupling, it has been created a new direction of development for this type of couplings.

Bibliography

[1] Automotive Engineering. Powertrain, Chassis System and Vehicle Body. Edited by David A. Crolla. Butterworth- Heinemann, 2009.

[2] Stone, R. BALL, J. Automotive Engineering Fundamentals. SAE International, 2004.

[3] Untaru, M. s.a. Calculul si constructia automobilelor. Editura didactica si pedagogica. Bucuresti, 1982.

[4] Preda I.Vehicle construction. Course notes.

[5] Ciolan, Gh. Preda, I. Dinamica autovehiculelor partea I. Editura Universitatii Transilvania. Brasov, 2009.

[6] Untaru, M. s.a. Dinamica autovehiculelor pe roti. Editura didactica si pedagogica. Bucuresti, 1981.

[7] Preda I. Vehicle dynamics. Course notes.

[8] Velicu R. Machine elements. Course notes.

[9] Deliu, G. Mechanics for engineering students. Editura Albastra. Cluj-Napoca, 2002

[10] *** Schaeffler Technical Pocket Guide. Schaeffler Technologies GmbH & Co. KG, 2014

[11] Ulrich, F. a.o. Mechanical and Metal Trades Handbook. Europa Lehrmittel, 2012.

[12] Buzdugan, Gh. Rezistenta materialelor. Editura Tehinca. Bucuresti, 1980.

[13] Velicu, R. Organe de masini. Universitatea Transilvania din Brasov, 2003. [14] Jula, A. s.a. Organe de masini, vol I si II. Universitatea din Brasov, 1986, 1989.

[15] Preda, I. Model Based Algorithm for the Study of the Vehicle Suspension

[16] http://www.tanabe-usa.com/manufacturing_spring.asp

[17] http://www.hotcoilsprings.com/hot-winding-springs.html

[18] http://www.omdspa.it/english/topic.asp?name=H150

[19] http://www.lsptechnologies.com/shot-peening/

[20] https://www.tokaibane.com/hotcoiled-spring/about/manufacturing/

[21] https://www.youtube.com/watch?v=t4n0LW1Eoj8

[22] www.auto.unitbv.ro/moodle

[23] http://www.ott-jakob.com.pl/pdf/Safety_whole_Inet_e.pdf

[24] https://en.wikipedia.org/wiki/Double_wishbone_suspension

[25] https://en.wikipedia.org/wiki/Torsion_bar_suspension

[26] http://www.madehow.com/Volume-6/Springs.html

[27] http://www.technicaljournalsonline.com/ijaers/VOL%20III/IJAERS%20VOL%20III%20ISSUE%20I%20%20OCTBER%20DECEMBER%202013/395.pdf

Annex – Drawings and assembly drawing