Introduction

Every year the car traffic on the roads of Russia is steadily increasing. In such conditions, the design of vehicles that meets modern traffic safety requirements is of paramount importance.

Driving safety is greatly influenced by the steering design as the most important factor in the driver's interaction with the road. To improve the characteristics of the steering, different types of amplifiers are added to its design. In our country, power steering is used almost only on trucks and buses. Abroad, more and more passenger cars have power steering, including cars of medium and even small classes, since power steering has an undoubted advantage over conventional ones, and provides much greater comfort and safety.

1.1 Initial data for steering design

Chassis parameters depend on the type of body, the location of the engine and gearbox, the mass distribution of the vehicle and its external dimensions. In turn, the steering scheme and design depend both on the parameters of the entire vehicle and on the decisions made on the scheme and design of other chassis and drive elements. Steering layout and design are determined early in the vehicle design.

The basis for the choice of the control method and the steering layout diagram are the characteristics and design solutions adopted at the stage of preliminary design, such as: maximum travel speed, base dimensions, track dimensions, wheel arrangement, axle load distribution, minimum turning radius of the vehicle.

In our case, it is necessary to design the steering for a small class passenger car with a front transverse engine and front drive wheels.

Initial data for calculations:

To assess the forces and moments acting in the steering, information is also required on the main kinematic points of the front suspension, as well as the angles of the steering wheels. Typically, these data become determined as the synthesis of the kinematic suspension scheme is completed at the end of the assembly stage and are refined (corrected) at the stage of the car fine-tuning. For initial, approximate calculations, the data on the angles of the pivot axis and the size of the run-in shoulder are sufficient. In our case, these are:

It should be noted that the accepted value of the minimum turning radius of the vehicle, which characterizes its maneuverability, is, apparently, the minimum possible for front-wheel drive vehicles of this class. The limiting factor here is the maximum possible angle in the constant velocity joints, which are used to transfer torque from the power unit to the front wheels. Analysis of the data on the turning radius of small cars produced in the 70-80s shows that its value lies in the range of 4.8-5.6 m. Further reduction of this indicator is possible only through the use of all-wheel steering.

To estimate (calculate) the moment on the steering wheel and the forces acting in the steering, it is necessary to know the axle load. For front-wheel drive vehicles, the average axle weight distribution is (%):

1.2 Purpose of steering. Primary requirements

Steering is a set of devices that rotate the steering wheels of a car when the driver acts on the steering wheel. It consists of a steering gear and a steering gear. To facilitate turning the wheels, an amplifier can be integrated into the steering gear or drive. In addition, a shock absorber can be integrated into the steering system to improve driving comfort and safety.

The steering gear is designed to transfer power from the driver to the steering gear and to increase the torque applied to the steering wheel. It consists of a steering wheel, a steering shaft and a gearbox. The steering drive is used to transfer force from the steering mechanism (gearbox) to the steered wheels of the vehicle and to ensure the required ratio between the angles of their rotation. The shock absorber compensates for shock loads and prevents steering wobble.

The task of the steering is the most unambiguous transformation of the steering wheel angle into the wheel angle and the transmission of information about the vehicle movement state to the driver through the steering wheel. The steering structure must provide:

1) Ease of control, assessed by the effort on the steering wheel. For cars without an amplifier when driving, this effort is 50 ... 100 N, and with an amplifier 10 ... 20 N. According to the project OST 37.001 "Vehicle handling and stability. General technical requirements", which was put into effect in 1995, the effort on the steering wheel for vehicles of category M 1 and M 2 must not exceed the following values.

The standards for the effort on the steering wheel, given in the draft OST, correspond to the enacted UNECE Regulations No. 79;

2) Rolling of the steered wheels with minimal side slip and slip when turning the car. Failure to comply with this requirement leads to accelerated tire wear and a decrease in vehicle stability while driving;

3) Stabilization of the turned steered wheels, ensuring their return to the position corresponding to straight-line movement with the steering wheel released. According to the project OST 37.001.487, the return of the steering wheel to the neutral position should occur without hesitation. One transition of the steering wheel through the neutral position is allowed. This requirement is also aligned with UNECE Regulation No. 79;

4) Informativeness of the steering, which is ensured by its reactive action. According to OST 37.001.487.88, the effort on the steering wheel for a car of category M 1 should increase monotonically with an increase in lateral acceleration up to 4.5 m / s 2;

5) Prevention of transmission of shocks to the steering wheel when the steered wheels hit an obstacle;

6) Minimum joint clearances. Evaluated by the angle of free rotation of the steering wheel of a car standing on a dry, hard and level surface in a position corresponding to straight-line movement. According to GOST 21398-75, this gap should not exceed 15 0 with the presence of an amplifier and 5 0 - without an amplifier of steering;

7) Absence of self-oscillations of the steered wheels when the car is operating in any conditions and in any driving modes;

8) The angles of rotation of the steering wheel for cars of category M 1 must be within the limits established by table. :

In addition to these basic functional requirements, the steering must provide a good "road feel" which also depends on:

1) a sense of precision control;

2) smoothness of the steering;

3) efforts on the steering wheel in the zone of rectilinear movement;

4) the feeling of friction in the steering;

5) sensation of viscosity of the steering;

6) the accuracy of centering the steering wheel.

At the same time, depending on the speed of the vehicle, different characteristics are of greatest importance. In practice, at this stage of the design process, it is very difficult to create an optimal steering design that would provide a good "road feel". Usually this problem is solved empirically, based on the personal experience of designers. The final solution to this problem is provided at the stage of fine-tuning the car and its components.

Special requirements are imposed on the reliability of the steering, since when it is locked, when any of its parts is destroyed or loosened, the car becomes uncontrollable, and an accident is almost inevitable.

All the stated requirements are taken into account when formulating particular requirements for individual parts and steering elements. So, the requirements for the sensitivity of the car to steering and to the maximum effort on the steering wheel limit the steering gear ratio. To provide a "feeling of the road" and reduce steering effort, the forward efficiency of the steering mechanism should be minimal, but from the point of view of the information content of the steering and its viscosity, the inverse efficiency should be high enough. In turn, a large value of efficiency can be achieved by reducing friction losses in the suspension and steering joints, as well as in the steering mechanism.

To ensure the minimum slip of the steered wheels, the steering linkage must have certain kinematic parameters.

Steering rigidity is of great importance for vehicle handling. With an increase in rigidity, control accuracy improves, and steering response increases.

Steering friction plays both a positive and a negative role. Low friction worsens the rolling stability of the steered wheels, increases the level of their vibrations. High friction reduces steering efficiency, increases steering effort, and impairs road feel.

Steering clearances also play both positive and negative roles. On the one hand, if they are present, jamming of the steering control is excluded, friction is reduced due to the "shaking" of the nodes; on the other hand, the "transparency" of the steering control deteriorates, its speed deteriorates; Excessive steering clearances can lead to self-oscillation of the steered wheels.

Special requirements are imposed on the geometric dimensions of the steering wheel and its design. An increase in the diameter of the steering wheel leads to a decrease in the effort on the steering wheel, however, it complicates its layout in the passenger compartment, worsens ergonomic performance and visibility. At present, the steering wheel diameter for small passenger cars of general purpose is 350 ... 400 mm.

The steering gear must provide a minimum clearance in the middle position of the steering wheel (corresponding to the straight-line movement of the vehicle). In this position, the working surfaces of the parts of the steering mechanism are subject to the most intense wear, that is, the play of the steering wheel in the middle position increases faster than in the extreme ones. To prevent jamming in the extreme positions when adjusting the gaps, the steering gear engagement is performed with an increased gap in the extreme positions, which is achieved by constructive and technological measures. During operation, the difference in the meshing clearances in the middle and extreme positions decreases.

The steering gear should have a minimum number of adjustments.

To ensure the passive safety of the vehicle, the steering wheel shaft must bend or disengage in an accident; the steering column tube and its fasteners must not interfere with this process. These requirements are implemented in the automotive industry in the form of safety steering columns. The steering wheel must deform in an accident and absorb the energy transmitted to it. At the same time, it should not collapse, form fragments and sharp edges. Front wheel limiters on the swing arms or on the steering box should reduce stiffness even under heavy loads. This prevents kinking of the brake hoses, tire rubbing against the mudguard and damage to suspension and steering components.

car steering gear rack

1.3 Analysis of known steering structures. Justification

selection of rack and pinion control

The steering wheel, through its shaft, transmits to the steering mechanism the torque developed by the driver and converts it into tensile forces on the one hand, and compression forces on the other, which, through the side rods, act on the pivoting levers of the steering linkage. The latter are fixed on pivot pins and rotate them to the required angle. The turning takes place around the pivot axles.

Steering gears are divided into rotary and reciprocating output mechanisms. Three types of steering mechanisms are installed on passenger cars: "worm-double-ridged roller", "screw-nut with circulating balls" - with a rotary movement at the output, and "gear-rack" - with a rotational-translational movement.

The circulating ball screw nut steering gear is quite advanced, but also the most expensive of all steering gears. In the screw pair of these mechanisms, there is not sliding friction, but rolling friction. The nut, being at the same time a rack, is in engagement with the toothed sector. Due to the small angle of rotation of the sector, it is easy for such a mechanism to realize a variable gear ratio with its increase as the angle of rotation of the rudder increases by setting the sector with eccentricity or by using a variable pitch of the gearing. High efficiency, reliability, stability of characteristics under heavy loads, high wear resistance, the possibility of obtaining a gap-free connection have led to the practical exclusive use of these mechanisms on cars of large and upper classes, partly in the middle class.

On passenger cars of small and very small classes, steering mechanisms of the "worm-roller" and "gear-rack" type are used. With the dependent suspension of the front wheels, which is currently used only on vehicles with high and high cross-country ability, a steering mechanism is required only with a rotary movement at the output. In terms of the overwhelming number of indicators, the mechanisms of the "worm-roller" type are inferior to the "gear-rack" mechanism and due to the convenience of the layout on front-wheel drive cars, the latter mechanisms are extremely widely used.

The advantages of rack-and-pinion steering are:

· Simplicity of construction;

· Low manufacturing costs;

· Ease of movement due to high efficiency;

· Automatic elimination of gaps between the rack and pinion, as well as uniform own damping;

· Possibility of hinged attachment of lateral transverse rods directly to the steering rack;

· Low pliability of the steering and, as a consequence, its high speed;

· The small volume required to install this steering system (due to which it is installed on all front-wheel drive cars produced in Europe and Japan).

· Lack of a pendulum arm (including its supports) and medium thrust;

· High efficiency due to low friction both in the steering mechanism and in the steering gear due to the reduction of the number of joints.

The disadvantages include:

· Increased sensitivity to shocks due to low friction, high return efficiency;

· Increased load from efforts from side rods;

· Increased sensitivity to steering fluctuations;

· Limited length of side rods (when they are hinged to the ends of the steering rack);

· Dependence of the angle of rotation of the wheels on the travel of the gear rack;

· Increased efforts in the entire steering system due to sometimes too short pivoting levers of the steering linkage;

· Reduction of the gear ratio with an increase in the angle of rotation of the wheels, as a result of which maneuvering in the parking lot requires great efforts;

· The impossibility of using this steering in vehicles with dependent suspension of the front wheels.

The following types of rack and pinion steering are most widely used:

Type 1 - lateral arrangement of the gear (left or right, depending on the position of the steering wheel) when attaching the lateral rods to the ends of the rack;

Type 2 - the middle arrangement of the gear with the same fastening of the steering rods;

Type 3 - lateral arrangement of the gear when attaching the lateral rods to the middle of the gear rack;

Type 4 - economical short version: lateral arrangement of the pinion by attaching both side rods to one end of the rack.

The Type 1 rack and pinion steering is the simplest design and requires the least amount of space. Since the hinges for the side links are fixed at the ends of the toothed rack. The rail is loaded mainly by axial forces. The radial forces, which depend on the angles between the side rods and the axis of the rack, are small.

In almost all front-wheel drive vehicles with a transverse engine arrangement, the steering linkage pivots are directed backward. If, in this case, due to a change in the height of the external and internal hinges of the side rods, the required inclination during cornering is not achieved, then, both during the compression stroke and during the rebound, the convergence becomes negative. Avoiding unwanted toe changes is possible in a vehicle with a low steering gear and side links slightly longer than the lower wishbones. A more favorable case is the front position of the steering linkage, which is practically achievable only for cars of the classic layout. In this case, the pivoting arms of the steering linkage must be turned outward, the outer hinges of the side links go deep into the wheels, the side links can be made longer.

Rack and pinion steering type 2, in which the gear is mounted in the midplane of the vehicle, is used only on cars with a mid or rear engine position, since the middle engine location entails such a disadvantage as a large required volume for steering due to the need for "kink "steering shaft.

In the event that the steering gear must be positioned relatively high, it is inevitable that the side rods are attached to the middle of the rack when using the MacPherson suspension. A diagram illustrating the basics of choosing the length of the side rods for the MacPherson suspension is shown in Fig. 1. In such cases, the inner joints of these rods are attached in the midplane of the vehicle directly to the rail or a member associated with it. In this case, the design of the steering mechanism should prevent the torsion of the gear rack by the moments acting on it. This imposes special requirements on the guide rails and drivers, since if the gaps are too small in them, the steering will be very difficult (due to high friction), if too large, knocks occur. If the cross-section of the toothed rack is not circular, but Y-shaped, then additional measures to prevent the rack from twisting around the longitudinal axis can be omitted.

Figure: 1. Determination of the length of the lateral link.

The type 4 steering system, which is installed on Volkswagen passenger cars, is easy to move and inexpensive to manufacture. The disadvantages include increased loads of individual parts and the resulting decrease in rigidity.

To prevent bending / twisting caused by the bending moment, the toothed rack has a relatively large diameter of 26 mm.

In practice, the choice of the type of rack and pinion steering is made from layout considerations. In our case, due to the lack of space for placing the steering mechanism at the bottom, the upper position of the steering mechanism is adopted. This necessitates the use of steering types 3.4. To ensure the strength and rigidity of the structure, the overhead steering arrangement and type 3 steering are finally adopted.

Admittedly, such a steering arrangement is not the most successful one. The high position of the steering gear makes it more flexible due to the deflection of the suspension struts. In this case, the outer wheel bends towards the positive camber, the inner wheel - towards the negative one. As a result, the wheels additionally tilt in the direction where the lateral forces tend to tilt them when cornering.

Kinematic calculation of the steering drive.

The kinematic calculation consists in determining the steering angles of the steered wheels, finding the gear ratios of the steering mechanism, drive and control as a whole, choosing the parameters of the steering linkage, as well as coordinating the steering and suspension kinematics.

1.4 Determination of the parameters of the steering link

First, the maximum average steering angle required to move the vehicle with the minimum radius is calculated. According to the diagram shown in Fig. 2.

(1)

Figure: 2.Schema turning the car with absolutely rigid wheels.

Figure: 3.Schema turning a car with flexible wheels.

In order for the steered rigid wheels to roll when turning without slipping, their instantaneous turning center must lie at the intersection of the axes of rotation of all wheels. In this case, the outer q n and inner q n angles of rotation of the wheels are related by the dependence:

(2)

where l 0 is the distance between the points of intersection of the axes of the pins with the supporting surface. Since these points practically coincide for front-wheel drive cars with the centers of contact of the wheels with the road (which is due to the small roll-in shoulder and the longitudinal angle of inclination of the kingpin),

It is possible to provide such a dependence only with the help of a rather complex kinematic drive scheme, however, the steering linkage allows you to get as close to it as possible.

Due to the lateral compliance of the tires, the wheels roll with lateral forces under the influence of lateral forces. The turning diagram of a car with flexible wheels is shown in Fig. 3. For highly elastic tires, the shape of the trapezoid is brought closer to a rectangle in order to increase the efficiency of the outer, more loaded wheel. On some vehicles the trapezoid is designed in such a way that the wheels remain approximately parallel up to a steering angle of »10 0. But at large angles of rotation of the wheels, the curve of the actual angles of rotation again reaches the curve of the required angles according to Ackermann. This reduces tire wear when parking and cornering.

The selection of trapezium parameters begins with determining the angle of inclination of the side trapezoid levers. Currently, this angle is usually selected based on the design experience of previous models.

For the designed steering, we take l \u003d 84.19 0.

Next, the length of the trapezium pivot arm is determined. This length is taken as large as possible according to the layout conditions. Increasing the length of the swing arm reduces the forces acting in the steering, as a result, increases the durability and reliability of the steering, as well as reduces its pliability.

In our case, the length of the pivoting arm is taken equal to 135.5 mm.

Obviously, with an increase in the length of the swing arm, the rack travel required to achieve a given maximum steering angle of the steered wheels increases.

The required rail travel is determined graphically or by calculation. Also, the kinematics of the steering linkage is determined graphically or by calculation.

Figure: 4. Dependence of the average angle of rotation of the steered wheels on the rack travel

In fig. 4 shows a graph of the dependence of the average angle of rotation of the wheels on the rack travel. The data for plotting was obtained using the WKFB5M1 program, which is used in the general layout department and the chassis department and the brakes department of the UPSh DTR VAZ to calculate the kinematics of the MacPherson suspension and rack and pinion steering. According to the graph, we determine that to ensure the angle of rotation of the wheels q \u003d 34.32 0, the rail travel in one direction equal to 75.5 mm is required. Full rail travel l \u003d 151 mm.

In fig. 5 shows the dependence of the difference between the angles of rotation of the outer and inner wheels as a function of the angle of rotation of the inner wheel. It also shows the curve of the required change in the difference between the angles of rotation of the wheels calculated according to Ackerman.

The indicator used to assess the kinematics of the steering drive is the difference in the angles of rotation of the wheels at the angle of rotation of the inner wheel equal to 20 0:

1.5 Steering gear ratio

The general kinematic steering ratio, determined by the gear ratios of the mechanism U r.m. and drive U r.p. is equal to the ratio of the total angle of rotation of the steering wheel to the angle of rotation of the wheels from lock to lock:

(5)

Figure: 5. Dependence of the difference between the angles of rotation of the wheels on the angle of rotation of the inner wheel:

1-calculated by the Ackermann ratio

2-for the designed car

For passenger cars with mechanical steering q r.k. max \u003d 1080 0… 1440 0 (3… 4 turns of the steering wheel), in the presence of an amplifier q r.k. max \u003d 720 0… 1080 0 (2… 3 turns of the steering wheel).

Usually, the number of revolutions of the steering wheel is determined within these limits based on the results of calculating the gear-rack gearing. In our case, the calculations showed the optimal number of revolutions equal to 3.6 (1296 0).

Then the total gear ratio is:

(6)

It is known that

(7)

Since a steering mechanism with a constant gear ratio is adopted for the designed car, U r.m. constant for any steering angle:

The steering gear ratio is not constant and decreases with increasing steering angle, which adversely affects the effort on the steering wheel when parking.

The dependence of the kinematic gear ratio of the designed steering is shown in Fig. 6

Figure: 6. Dependence of the steering gear ratio on the steering angle.

There are two approaches to matching suspension and steering kinematics. According to the first, during the rebound and compression strokes of the suspension, the steering wheels should not turn; According to the second, more advanced one, the designer deliberately sets the law of changing the toe-in of the wheels during the suspension strokes to improve the vehicle's handling and reduce tire wear. According to the recommendations of the Porsche company, which are used at VAZ in the design, the toe-in of the wheels should increase during rebound and decrease during the compression of the suspension. The rate of toe change should be 3-4 minutes per centimeter of suspension travel.

This work is carried out by the specialists of the general layout department and the synthesis of the suspension and steering kinematics is included, as a result of which the coordinates of the characteristic kinematic points are determined.

1.7 Calculation of the gearing parameters of the gear-rack mechanism

The calculation of the parameters of the gearing of the gear-rack transmission has a number of features. Since this transmission is low-speed and also backlash-free, special requirements are imposed on the profile of the gear and rack teeth.

Initial data for calculations:

1. Module according to nomograms, usually from the standard series (1.75; 1.9; 2.0; ...) depending on the rack travel and the number of revolutions of the steering wheel: m 1 \u003d 1.9

2. Number of gear teeth z 1. Also selected according to nomograms. For rack and pinion steering mechanisms usually lies in the range of 6 ... 9. z 1 \u003d 7

3. The angle of the original contour a and.sh. \u003d 20 0

4. The angle of inclination of the pinion shaft axis to the longitudinal axis of the rack d \u003d 0 0.

5. The angle of inclination of the gear tooth b.

The smallest slip and, consequently, the highest efficiency is provided at b \u003d 0 0. in this case, axial loads do not act on the bearings of the pinion shaft.

Helical gearing is adopted when it is necessary to ensure increased strength, as well as for mechanisms with a variable gear ratio - to ensure smooth operation.

We accept b \u003d 15 0 50 ".

6. Center distance a. It is usually taken as the minimum possible in terms of strength, which provides a compact design, reduces the weight of the steering mechanism and provides a good layout. a \u003d 14.5 mm

7. Rail diameter d. To ensure the strength of the mechanism due to the length of the tooth, we take d \u003d 26 mm.

8. The rail travel l p \u003d 151 mm.

9. Coefficient of the radial clearance of the gear C 1 \u003d 0.25 mm.

10. Ratio of the tooth head of the gear making tool

11. Coefficient of the radial clearance of the rail C 2 \u003d 0.25 mm.

12. Ratio of the tooth head of the tool for making a rack

Calculation of gear parameters:

1. The coefficient of displacement of the original contour is minimal (determined from the condition of the maximum profile overlap)

2. The minimum diameter of the tooth stem.

3. Diameter of the main circle

(10)

4. The diameter of the starting circle

(11)

5. Ratio of tooth head height

(12)

6. Angle of engagement (face angle) during manufacture

7. The maximum coefficient of displacement of the original contour x 1 max is determined from the condition that the thickness of the tooth head is equal to 0.4m 1. The calculation requires the diameter of the circumference of the tooth head d a 1. a preliminary calculation of the diameter of the tooth head is carried out according to the formula:

, (see Fig. 7.) (14)

The angle a SK is taken equal to 50 0, and then it is corrected by the operational method according to the formula:

(15)

where - correction to the angle a SK (rad);

(17)

Sufficient accuracy in calculating a SK is achieved after 4 operations

Then

(18)

8. Coefficient of displacement of the original contour x 1 is selected within x 1 min

9. Diameter of the circumference of the gear tooth head d a 1 with the selected x 1:

d a 1 \u003d 2m 1 (h * 01 + x 1) + d 01 \u003d 19.87mm (19)

10.The diameter of the circumference of the foot of the gear tooth

11. The diameter of the active circle of the gear tooth foot d n 1 is calculated depending on the sign of B:

d n 1 \u003d d B 1 for B £ Ф (21)

at B\u003e Ф (22)

where (23);

h * a2 - ratio of the rack tooth head

d n 1 \u003d 13,155 mm

Gear tooth height

(24)

12. Angle a SK with the accepted coefficient of displacement of the original contour x 1:

(25)

13. The proportional overlap in the end section e a is calculated depending on A:

(27) at A<Ф

where A \u003d a-r Na 2 -0.5d B 1 cosa wt is the distance between the active line of the rack tooth head and the main circle;

r Na 2 - distance from the staff axis to the active line of the tooth head

14. Axial overlap in the end section

(28)

where b 2 is the average width of the rack tooth

15. End module

(29)

16. Radial clearance of gear

C 1 \u003d m n C 1 * \u003d 0.475 mm (30)

17. Basic step

P b \u003d pm n cosa 01 \u003d 5.609 mm (31)

18. Coefficient of displacement of the original contour in the end section

x f1 \u003d x n1 × cosb 1 \u003d 0.981 (32)

19. Thickness of the tooth on the base circle in the end section

S bt1 \u003d (2 х 1 tga 0 + 0.5p) cosa wt m t + d B1 × inva wt \u003d 4.488210mm (33)

inv a wt \u003d tga wt –a wt / 180 \u003d 0.01659 (34)

20. Thickness of the gear tooth head

Pinion diameter at the end of the rack

for d a 1 -d y\u003e 0 for d a 1 -d y £ Ф d a 1 \u003d d y

where r Na 2 is the distance from the rod axis to the active line of the tooth head

21. Measured number of gear teeth

(37)

rounded down, where b B \u003d arcsin (cosa 0 × sinb 01) is the angle of inclination of the tooth along the main circle;

P l \u003d pm n cosa 01 - main step

22. The length of the common normal

W \u003d (z "-1) P b + S bt1 cosb B \u003d 9.95mm (38)

23. Minimum active gear width

1.8 Calculation of rail parameters

1. Angle of inclination of the tooth of the rack

b 02 \u003d d-b 01 \u003d -15 0 50 "(40)

2. Ratio tooth head ratio

h * a2 \u003d h * ap01 -C * 2 \u003d 1.25 (41)

3. Radial clearance of the rack

C 2 \u003d m n C * 2 \u003d 0.475 (42)

4. Distance from the axis of the rack to the centerline of the tooth

r 2 \u003d a-0.5d 01 -m n x 1 \u003d 5.65 mm (43)

5. Distance from the axis of the staff to the line of the tooth stem

r f2 \u003d r 2 -m n h * ap02 \u003d 4.09 mm (44)

6. Distance from the axis of the staff to the active line of the tooth head

r Na2 \u003d r 2 + m n h * ap01 -m n C * 2 \u003d 8.025mm (45)

7. Distance from the axis of the rack to the line of the tooth head of the rack

r a 2 \u003d r Na 2 + 0.1 \u003d 8.125 (46)

8. Average tooth width of the rack

9. Distance from the axis of the staff to the active line of the tooth stem

r N2 \u003d a-0.5d a1 cos (a SK -a wt) \u003d 5.78 mm (48)

10. Height of the rack tooth head

h a2 \u003d r a2 -r 2 \u003d 2.475 mm (49)

11. Height of the leg of the rack tooth

h f2 \u003d r 2 -r f2 \u003d 1.558mm (50)

12. Height of the rack tooth

h 2 \u003d h a 2 - h f 2 \u003d 4.033 mm (51)

13. End step

(52)

14. Thickness of the rack tooth at the leg

S fn2 \u003d 2 (r 2 - r f2) tga 0 + 0.5pm n \u003d 4.119 mm (53)

15. Width of the depression at the leg

S ef2 \u003d pm n - S fn2 \u003d 1.85 mm (54)

16. Thickness of the rack tooth head

S an2 \u003d 0.5 pm n - (r Na2 + 0.1- r 2) 2tga 0 \u003d 1.183 mm (55)

17. Radius of the base of the leg of the rack tooth

P f2 \u003d 0.5 S ef2 × tg (45 0 + 0.5d 0) \u003d 1.32 mm (56)

18. Minimum number of rack teeth z 2 min:

where l p is the rail travel

Loss of length (difference between total engagement and rail travel) (58);

(59)

l 1 \u003d a-r a2 (60)

(62)

(63)

19. Diameter of the measuring roller theoretical

round up to the existing d 1 \u003d 4.5 mm

20. Measured dimension from the edge of the rail

21. Measured diameter from the rail axis

22. Measured diameter to the tooth head

23. Measured diameter to the root of the tooth

Chassis parameters depend on the type of body, the location of the engine and gearbox, the mass distribution of the vehicle and its external dimensions. In turn, the steering scheme and design depend both on the parameters of the vehicle as a whole and on the decisions made on the scheme and design of other chassis and drive elements. The steering layout and design are determined early in the vehicle design phase.

The basis for the choice of the control method and the layout of the steering circuit are the characteristics and design solutions adopted at the stage of preliminary design: maximum speed, base size, wheel formula, axle load distribution, minimum turning radius of the vehicle, etc.

The steering of a VAZ-2110 car consists of a rack-and-pinion steering mechanism and a steering drive. The design presented in the graphic part of this diploma project is a rack and pinion steering gear assembly with rods, as well as working drawings of its parts.

Rack and pinion steering mechanisms are more common, since they have a low mass, high efficiency and increased rigidity, they are well combined with hydraulic amplifiers, which led to their use in passenger cars with a front engine, for example, on the VAZ-2110, steering is used due to the fact that that this car model has a maximum steering axle load of up to 24 kN.

The steering diagram of a VAZ-2110 car is shown in Fig. 8. In this figure:

1 - thrust tip head;

2 - ball joint;

3 - swivel levers;

5 - tubular rod;

6 - horizontal rods;

8 - fastening rod;

12 - connecting plate;

13 - lock plate;

14 - rubber-metal hinge;

15 - sealing rings;

16 - bushing;

17 - rail;

18 - crankcase;

19 - clamp;

20 - elastic coupling;

21 - steering rods;

22 - damping element;

23 - steering wheel;

24 - deep groove ball bearing;

26 - steering column;

27 - bracket;

28 - protective cap;

29 - roller bearing;

30 - drive gear;

31 - ball bearing;

32 - retaining ring;

33 - protective washer;

34 - sealing rings;

35 - nut;

36 - anther;

37 - rubber ring;

38 - retaining ring;

39 - cermet stop;

40 - spring;

44 - nut.

Figure 9 shows a rack and pinion steering gear assembly.

This design includes:

1 - protective cap;

2 - steering gear housing;

3 - steering rack;

4 - drive gear;

5 - steering rod;

6 - spacer sleeve that limits the rail travel;

7 - the bolt of fastening of the steering link, tighten with the moments of 7.8 ± 0.8 kgf × m and lock them by bending the edges of the locking plate on the verge of the bolts;

8 - connecting plate;

9 - persistent sleeve;

10 - support of the steering mechanism, tightly fitting to the cover;

11 - support sleeve of the rail;

12 - protective cover, installed so that its right end is at a distance of 28.5 -0.5 mm from the end of the pipe, and secured with clamps;

13 - clamp;

14 - thrust ring of the rack, which limits the rack travel;

15 - sealing ring of the rail stop;

16 - nut;

17 - rail stop;

18 - roller bearing;

19 - ball bearing;

The set screw is loaded by a radial force F r \u003d 985 H and F L 1 \u003d 1817.6 H.

Thread M32 x 1.5

Material:

Grub screw GD - Z and Al 4

Bushing CDAl 98 Cu 3

Carrying thread length 5 mm.

Contact voltage

Material for all force-transmitting parts, such as steering link arms, swing arms, transverse links, ball joints, etc., must have a sufficiently large elongation. When overloaded, these parts should deform plastically, but not collapse. Parts made of materials with low elongation, such as cast iron or aluminum, must be correspondingly thicker. When the steering is locked, when any of its parts is destroyed or loosened, the car becomes uncontrollable, and an accident is almost inevitable. This is why the reliability of all parts is important.

6. Ilarionov V.A., Morin N.M., Sergeev N.M. Theory and design of the car. Moscow: Mechanical Engineering, 1972

7. Loginov M.I. Car steering. Moscow: Mechanical Engineering, 1972

8. Lukin P.P., Gaparyants G.A., Rodionov V.F. Car design and calculation. Moscow: Mechanical Engineering, 1984

9. Labor protection in mechanical engineering. M.: mechanical engineering, 1983

10. Labor protection at road transport enterprises. Moscow: Transport, 1985

11. Raimpel J. Car chassis. Moscow: Mechanical Engineering, 1987

12. Tchaikovsky I.P., Solomatin P.A. Steering controls of cars. M. Mechanical Engineering, 1987

A. A. Enaev

Cars.

Design and calculation

steering

Study guide

Bratsk 2004

2. PURPOSE, REQUIREMENTS AND CLASSIFICATION ... 3. SELECTING THE METHOD OF TURNING CARS ……… 4. SELECTION OF THE STEERING SYSTEM ……………. 5. STEERING GEARS ………………………………… .. 5.1. Purpose, requirements, classification …………… ... 5.2. Estimated parameters of the steering gear ………… .. 5.3. Selection of the type of steering mechanism ………………………. 5.4. Materials used for the manufacture of steering gears ………………………………………………… ... 6. STEERING GEARS ………………………………………. 6.1. Purpose, requirements, classification …………… ... 6.2. Estimated parameters of the steering gear …………… .. 6.3. Selection of the type of steering drive …………………………. 6.4. Materials used for the manufacture of steering gears ……………………………………………………… 7. POWER STEERING ……………… .. 7.1. Purpose, requirements, classification …………… ... 7.2. Estimated parameters of the power steering ……………………………………………………………. 7.3. The choice of the layout of the amplifiers ... ... ... ... ... ... 7.4. Amplifier pumps …………………………………… ... 7.5. Materials used for the manufacture of pump amplifiers ………………………………………………… ... 8. CALCULATION OF THE STEERING CONTROL …………………… ... 8.1. Kinematic calculation of the steering drive ……………. 8.2. Steering gear ratio ……………. 9. POWER CALCULATION OF THE STEERING CONTROL ... ... ... ... 9.1. Steering wheel force ……………………………… 9.2. The force developed by the amplifier cylinder ………… .. 9.3. Effort on wheels during braking ………………… ... 9.4. Efforts on the lateral and longitudinal rods …………… 10. HYDRAULIC CALCULATION OF THE AMPLIFIER …………… 11. STRENGTH OF THE STEERING CONTROL .. 11.1. Calculation of steering mechanisms ………………………… ... 11.2. Steering drive calculations ……………………………

Design and calculation of steering controls is one of the components of the course project in the discipline "Automobiles".

At the first stage of course design, it is necessary to perform a traction calculation and investigate the operational properties of a car using the guidelines “Cars. General Provisions. Traction calculation "and then proceed, in accordance with the assignment, with the design and calculation of the unit or chassis system of the vehicle.

When designing and calculating steering controls, it is necessary to select the recommended literature, carefully read this manual. The sequence of work on the design and calculation of steering controls is as follows:

1. Choose the way of turning the car, steering scheme, type of steering mechanism, amplifier layout scheme (if necessary).

2. Perform kinematic calculation, power calculation, hydraulic calculation of the amplifier (if the steering is provided for the installation of the amplifier).

3. Select the dimensions of the parts and perform the strength calculation.

This training manual details how to perform all these types of work.

2. PURPOSE, REQUIREMENTS AND CLASSIFICATION

Steering - a set of devices that serve to turn the steered wheels of a car when the driver acts on the steering wheel and consists of a steering mechanism and a drive (Fig. 1).

The steering gear is the part of the steering from the steering wheel to the steering bipod, and the steering gear includes parts from the steering bipod to the stub axle.

Figure: 1. Steering diagram:

1 - steering wheel; 2 - steering shaft; 3 - steering column; 4 - reducer; 5 - steering bipod; 6 - longitudinal steering rod; 7 - pivot pin; 8 - pivot pivot lever; 9 - side lever; 10 - transverse thrust

The following requirements are imposed on the steering:

1) ensuring high maneuverability of vehicles, in which sharp and fast turns are possible in relatively limited areas;

2) ease of control, assessed by the amount of force applied to the steering wheel.

For cars without an amplifier when driving, this effort is 50 ... 100 N, and with an amplifier - 10 ... 20 N. For trucks, the effort on the steering wheel is regulated: 250 ... 500 N - for steering without an amplifier; 120 N - for power steering;

3) rolling of steered wheels with minimal lateral slip and slip when turning the car;

4) the accuracy of the tracking action, primarily kinematic, in which any given position of the steering wheel will correspond to a well-defined previously calculated curvature of the turn;

Loads and stresses acting in the steering parts can be calculated by setting the maximum force on the steering wheel or by determining this force by the maximum resistance to turning the steering wheels of the car on the spot (which is more appropriate). These loads are static.

AT steering gear calculate the steering wheel, steering shaft and steering gear.

Maximum effort on steering wheel for steering systems without amplifiers - \u003d 400 N; for cars with amplifiers -
\u003d 800 N.

When calculating the maximum effort on the steering wheel based on the maximum resistance to turning the steered wheels in place, the moment of resistance to turning can be determined by the empirical relationship:

, (13.12)

where –The coefficient of adhesion when turning the steered wheel in place;
- wheel load;
–The air pressure in the tire.

The effort on the steering wheel for turning on the spot is calculated by the formula:

, (13.13)

where
- angular steering ratio;
–The radius of the steering wheel;
- Steering efficiency.

For a given or found effort on the steering wheel, the loads and stresses in the steering parts are calculated.

Spokes the steering wheel is designed to be bent, assuming that the steering wheel force is equally distributed between the spokes. The bending stresses of the spokes are determined by the formula:

, (13.14)

where
–The length of the spoke;
- spoke diameter;
- the number of spokes.

Steering shaft usually tubular. The shaft is torsionally loaded with a moment:

. (13.15)

The torsional stress of the tubular shaft is calculated by the formula:

, (13.16)

where
,
- outer and inner diameters of the shaft, respectively.

Permissible torsion stresses of the steering shaft - [
] \u003d 100 MPa.

The steering shaft is also checked for torsional rigidity:

, (13.17)

where
–Shaft length;
- modulus of elasticity of the 2nd kind.

Allowable angle of twisting - [
] \u003d 5 ÷ 8 ° per meter of shaft length.

AT worm and roller steering gear the globoid worm and the roller are calculated for compression, the contact stresses in engagement at which are determined by the formula:

, (13.18)

where –Axial force acting on the worm;
- contact area of \u200b\u200bone roller ridge with a worm; –Number of roller ridges.

The axial force acting on the worm is calculated by the formula:

, (13.19)

where - the initial radius of the worm in the smallest section;
- the angle of the helix of the worm.

The contact area of \u200b\u200bone roller ridge with a worm can be determined by the formula:

where and - the radius of engagement of the roller and the worm, respectively; and
- angles of engagement of the roller and the worm.

Allowable compression stresses - [
] \u003d 2500 ÷ 3500 MPa.

AT propeller driven gear the pair "screw - ball nut" is checked for compression taking into account the radial load on one ball:

, (13.21)

where
– number of working turns;
– the number of balls on one turn (when the groove is completely filled);
– the angle of contact of the balls with the grooves.

The strength of the ball is determined by the contact stresses calculated by the formula:

, (13.22)

where
– coefficient of curvature of contacting surfaces; – modulus of elasticity of the 1st kind;
and
– ball and groove diameters respectively.

Permissible contact voltages [
] \u003d 2500 ÷ 3500 MPa.

In the “rack - sector” pair, the teeth are calculated for bending and contact stresses similar to cylindrical gearing. In this case, the circumferential force on the teeth of the sector (in the absence or inoperative amplifier) \u200b\u200bis determined by the formula:

, (13.23)

where Is the radius of the initial circle of the sector.

Allowable voltages - [
] \u003d 300 ÷ 400 MPa; [
] \u003d 1500 MPa.

Rack and pinion steering gear calculated in the same way.

AT steering gear calculate the steering bipod shaft, steering bipod, steering bipod pin, longitudinal and transverse steering rods, swivel arm and levers of steering knuckles (pivots).

Steering arm shaft count on torsion.

In the absence of an amplifier, the bipod shaft voltage is determined by the formula:

, (13.24)

where - bipod shaft diameter.

Allowable voltages - [
] \u003d 300 ÷ 350 MPa.

Bipod calculation bend and torsion in a dangerous section AND-AND.

In the absence of an amplifier, the maximum force acting on the ball pin from the longitudinal steering rod is calculated by the formula:

, (13.25)

where - the distance between the centers of the steering arm heads.

The bipod bending stresses are determined by the formula:

, (13.26)

where - bipod bend shoulder; a and b - the dimensions of the bipod section.

The torsional stresses of the bipod are determined by the formula:

, (13.27)

where –The torsion shoulder.

Allowable voltage [
] \u003d 150 ÷ \u200b\u200b200 MPa; [
] \u003d 60 ÷ 80 MPa.

Bipod ball pin rely on bending and shearing in a dangerous section B-B and crushing between the longitudinal tie rod crumbs.

The bipod finger bending stress is calculated by the formula:

, (13.28)

where e - finger bend shoulder;
- the diameter of the finger in the dangerous section.

Finger shear stress is determined by the formula:

. (13.29)

Finger crush stresses are calculated using the formula:

, (13.30)

where - the diameter of the ball head of the finger.

Allowable voltages - [
] \u003d 300 ÷ 400 MPa; [
] \u003d 25 ÷ 35 MPa; [
] \u003d 25 ÷ 35 MPa.

Calculation of ball pins of longitudinal and transverse steering rods is carried out similarly to the calculation of the ball pin of the steering arm, taking into account the acting loads on each pin.

Longitudinal steering rod expect compression and buckling.

H compression stresses are determined by the formula:

, (13.31)

where
Is the cross-sectional area of \u200b\u200bthe thrust.

During buckling, critical stresses arise in the thrust, which are calculated by the formula:

, (13.32)

where - modulus of elasticity of the 1st kind; J - moment of inertia of the tubular section; - the length of the pull at the centers of the ball pins.

The thrust stability margin can be determined by the formula:

. (13.33)

The traction stability margin should be -
\u003d 1.5 ÷ 2.5.

Transverse tie rod loaded with force:

, (13.34)

where
and Are the active lengths of the swing arm and the swing arm, respectively.

The tie rod is designed for compression and buckling in the same way as the steering rod.

Swing arm count on bending and torsion.

. (13.35)

. (13.36)

Allowable voltages - [
] \u003d 150 ÷ \u200b\u200b200 MPa; [
] \u003d 60 ÷ 80 MPa.

Steering knuckle levers also count on bending and torsion.

Bending stresses are determined by the formula:

. (13.37)

Torsional stresses are calculated by the formula:

. (13.38)

Thus, in the absence of an amplifier, the strength calculation of steering parts is based on the maximum effort on the steering wheel. In the presence of an amplifier, the steering drive parts located between the amplifier and the steered wheels are, in addition, loaded with the force developed by the amplifier, which must be taken into account when carrying out calculations.

Amplifier calculation usually includes the following steps:

choice of amplifier type and layout;

static calculation - determination of forces and displacements, dimensions of the hydraulic cylinder and switchgear, centering springs and areas of reaction chambers;

dynamic calculation - determination of the amplifier turn-on time, analysis of oscillations and stability of the amplifier;

hydraulic calculation - determination of pump performance, pipe diameters, etc.

As reference loads, acting on the steering parts, we can take the loads arising from collisions of the steered wheels on road irregularities, as well as loads arising in the steering drive, for example, when braking due to unequal braking forces on the steered wheels or at rupture. tires of one of the steered wheels.

These additional calculations allow for a more complete assessment of the strength characteristics of steering parts.

Steering drive,which is a system of rods and levers, serves to transfer the force from the bipod to the pivot pins and to implement a given relationship between the angles of rotation of the steered wheels. When designing steering controls, the kinetic and power calculation of the steering drive and the strength calculation of the steering units and parts are performed.

The main task of the kinematic calculation of the steering drive is to determine the angles of rotation of the steered wheels, to find the gear ratios of the steering mechanism, drive and control as a whole, to select the parameters of the steering linkage and to coordinate the steering and suspension kinematics. Based on the trolleybus turning geometry (Fig. 50), provided that the steered front wheels roll without slipping and their instantaneous turning center lies at the intersection of the axes of rotation of all wheels, the outer and inner angles of rotationwheels are related by dependency:

, (4)

where is the distance between the points of intersection of the axes of the pins with the supporting surface.

Figure 50. Scheme of trolleybus turning without taking into account the lateral elasticity of the tires.

From the obtained expression (4) it follows that the difference between the cotangents of the angles of rotation of the external and internal steered wheels should always be constant, and the instantaneous center of rotation of the trolleybus (point 0) should lie on the continuation of the uncontrolled axis.

Only if these theoretical conditions are met, the weight of the trolleybus wheel on the bend will move without slipping, i.e. have clean rolling. The steering linkage is required to provide the ratios arising from the steering geometry between the steering angles of the steered wheels.

The parameters of the steering linkage are the pivot width (Fig. 51), the distance pbetween the centers of the ball joints of the trapezoid levers; length tand angle θ tilt of the pivot levers. The selection of the trapezium parameters with laterally rigid steered wheels begins with determining the angle θ tilt of the trapezoid levers. They are positioned so that and -(0.7...0.8,)L with the rear location of the transverse link. Angle θ can be found for maximum theoretical angles and according to the formula:

or according to the graphs shown in (Figure 7b). Angle value θ \u003d 66 ... 74 °, and the ratio of the length of the levers to the length of the transverse rod t / n \u003d0.12 .... 0.16. Length m take the largest possible layout. Then

.

Figure 51. Diagram of the steering linkage and dependence a / L from l 0 / L 1-3: at m / n equal to 0.12, respectively; 0.14; 0.16

The total kinematic steering ratio, determined by the gear ratios of the mechanism U mand drive U pcequal to the ratio of the total steering wheel angle to the wheel angle from lock to lock

.

For normal operation of the steering drive, the maximum value of the angles a, and a, is within
... For trolleybuses, the total number of turns of the steering wheel when the steered wheels are turned 40 ° (± 20 °) from the neutral position should not exceed 3.5 ( = 1260 о) without taking into account the angle of free rotation of the steering wheel, which corresponds to .

The schematic layout of the steering drive is performed to determine the size and location in space of the bipod, rods and levers, as well as the drive gear ratio. At the same time, they strive to ensure the simultaneous symmetry of the extreme positions of the bipod relative to its neutral position, as well as the equality of the kinematic gear ratios of the drive when turning the wheels both to the right and to the left. If the angles between the bipod and the longitudinal rod, as well as between the rod and the swing arm in its extreme position, are approximately the same, then these conditions are met.

In the power calculation, the forces are determined: required to turn the steered wheels in place, developed by the amplifier cylinder; on the steering wheel when the amplifier is working and not working; on the steering wheel from the side of the reactive elements of the distributor; on wheels when braking; on individual steering parts.

Force Frequired to turn the steered wheels on the horizontal surface of the trolleybus is found based on the total moment M Σon the axles of the steered wheels:

where M f–The moment of resistance to rolling of the steered wheels when turning around the pivots; M φ–The moment of resistance to tire deformation and friction in contact with the supporting surface as a result of tire slippage; M β, M φ- moments due to the lateral and longitudinal tilt of the pivots (Fig. 8).

Figure 52. Calculation of the moment of resistance to turning the wheel.

The moment of resistance to rolling of the steered wheels when turning around the pivots is determined by the relationship:

,

where f- rolling resistance coefficient; G 1- axle load transmitted by the steered wheels; - radius of wheel rolling around the pivot axis: \u003d 0.06 ... 0.08 m; l–The length of the journal; r 0- estimated wheel radius; λ - wheel camber angle; β - the angle of inclination of the king pin.

The moment of resistance of tire deformation and friction in contact with the supporting surface due to tire slippage are determined by the relationship:

,

where is the shoulder of the sliding friction force relative to the center of the tire indentation.

If we assume that the pressure is evenly distributed over the area of \u200b\u200bthe print,

,

where is the free radius of the wheel. In the case when.

When calculating the coefficient of adhesion to the supporting surface, the maximum φ= 0.8.

The moments due to the lateral and longitudinal inclination of the pivots are equal:

where is the average angle of rotation of the wheel; ; γ - the angle of inclination of the king pin back.

Force on the steering wheel rim

,

where is the radius of the steering wheel; η - Steering efficiency: η= 0.7…0.85.

Calculation of steering elements

The loads in the steering and steering components are determined based on the following two design cases˸

According to a given design effort on the steering wheel;

For maximum steering resistance in place.

When the vehicle is driven on uneven roads or when braking with different grip coefficients under the steer wheels, a number of steering components absorb dynamic loads that limit the strength and reliability of the steering. The dynamic impact is taken into account by introducing a dynamic factor to q \u003d 1.5 ... 3.0.

Estimated effort on the steering wheel for passenger cars P PK \u003d 700 N. To determine the effort on the steering wheel from the maximum resistance to turning the steered wheels in place 166 Steering, it is necessary to calculate the moment of resistance to turning according to the following empirical formula

M c \u003d (2p about / 3) V O b k / r w ,

where p about - coefficient of adhesion when turning the wheel in place ((p about \u003d 0.9 ... 1.0), G k - load on the driven wheel, p w - air pressure in the tire.

Effort on the steering wheel for turning in place

Р w \u003d Mc / (u a R PK nPp y),

where u a is the angular gear ratio.

If the calculated value of the steering wheel effort is greater than the above conditional design effort, then the vehicle requires the installation of a steering amplifier. Steering shaft. In most designs, ᴇᴦο is hollow. The steering shaft is loaded with torque

M PK \u003d P PK R PK .

Hollow shaft torsional stress

m \u003d M PK D /. (8.4)

Allowable stress [t] \u003d 100 MPa.

The angle of twist of the steering shaft is also checked, which is allowed within 5 ... 8 ° per one meter of the shaft length.

Steering gear. For a mechanism that includes a globoid worm and a roller, the contact stress in the engagement is determined

o \u003d Px / (Fn), (8.5)

P x - axial force perceived by the worm; F is the contact area of \u200b\u200bone roller ridge with the worm (the sum of the areas of the two segments, Fig. 8.4), and is the number of roller ridge.

Axial force

Px \u003d Mrk / (r wo tgP),

The material of the worm is cyanized steel ZOKH, 35X, 40X, ZOKHN; roller material - case-hardened steel 12ХНЗА, 15ХН.

Allowable voltage [a] \u003d 7 ... 8MPa.

For a screw-rack mechanism in the "screw-ball nut" link, the conditional radial load P 0 per ball is determined

P w \u003d 5P x / (mz COs - $ con),

where m is the number of working turns, z is the number of balls on one turn, 8 fin is the angle of contact of the balls with the grooves (d fin \u003d 45 o).

Contact voltage determining the strength of the ball

where E is the modulus of elasticity, d m is the diameter of the ball, d k is the diameter of the groove, k kr is the coefficient depending on

curvature of the contacting surfaces (k cr \u003d 0.6 ... 0.8).

Allowable stress [a (W] \u003d 2500..3500 MPa based on the diameter of the ball According to GOST 3722-81, the breaking load acting on one ball must be determined.

Calculation of steering elements - concept and types. Classification and features of the category "Calculation of steering elements" 2015, 2017-2018.