Steering with integral steering - diploma

Contents

Contents

Introduction

1 Information and patent review of materials on the topic of the diploma project

2 Justification and selection of design of designed units

3 Selection of basic parameters and calculation of designed units

3.1 Selection of main steering parameters

3.2 Kinematic and power steering calculation

3.3 Static and hydraulic calculations of steering hydraulic booster

3.4 Strength calculation of steering parts

3.5 Selection of main parameters of front axle

3.6 Calculation of strength of front axle parts

4 Process part of the project

5 Economic part of the project

6 Occupational safety

Conclusion

List of sources used

Application

Introduction

The diploma project developed steering with an integral steering mechanism and a rotary distributor of a hydraulic amplifier, as well as a beam of a front non-driving bridge for a city bus of the MAZ103 type.

Steering control of all MAZ buses is developed on the basis of units and assemblies of MAZ production cars.

On the low-floor buses MAZ103 and MAZ103C, the driver's force is transmitted through a steering wheel, a steering column adjustable in height and angle, an upper cardan shaft, an angular reduction gear, a lower cardan shaft, a steering mechanism with an integrated power steering distributor, a longitudinal steering rod, a pendulum lever and an intermediate steering rod to the right steering wheel. The right controlled wheel is connected to the left transverse steering rod.

On MAZ104, MAZ-104C, MAZ105 and MAZ152 buses, the driver's force is transmitted through a steering wheel, a steering column adjustable in height and angle, an upper cardan shaft, an angular reduction gear, a lower cardan shaft, a steering mechanism with an integrated power steering distributor, an intermediate steering rod, a pendulum lever and a longitudinal steering rod to the left steering wheel. The right controlled wheel is connected to the left transverse steering rod.

In both cases, the power cylinder is fixed by one end on the bus body, and with a rod through a tip unified with the tip of the steering rods is connected to the pendulum lever (the project provides for the creation of an integral steering mechanism, that is, with a combined power cylinder).

The distributor of the rudder hydraulic booster - spool type (the design provides for a rotor type distributor) is built into the steering mechanism, and serves to control the flow of working fluid from the pump to the cavities of the power cylinder, as well as to communicate the cavities of the power cylinder with each other when the working fluid is suddenly stopped from the hydraulic booster pump (this is necessary for the ability to control the bus only by the driver).

The steering mechanism is equipped with a pressure unloading valve in extreme positions of the nozzle, which increases the overhaul life of the steering rod tips. Valve is connected by discharge pipe with pressure line at distributor inlet. When the shaft-sector reaches the limit turning angles, the valve opens the channel connected to the pressure line and the internal volume of the steering mechanism, which constantly communicates with the drain of the hydraulic system. The steering gear parts are lubricated with oil coming from the hydraulic system of the steering power (the steering gear case is completely filled in).

On all buses, steering tips unified with MAZ6422 cars are used. Tips of additional and transverse steering rods have right and left threads for adjustment of length without removal of tips. Tips on rods are fixed by tightening bolts

The tip of the steering rod consists of a body, in which a ball pin is installed between the crumbs. The breadcrumbs are pressed against the spherical head of the pin by a spring. The spring is pre-compressed when the plug is tightened. After adjustment, the plug is locked by the cover when the bolts are tightened. On the side of the cone part of the pin, the hinge is sealed with a seal. To lubricate the hinge, an oiler is screwed into the tip housing.

Depending on the engine installed on the bus, various configurations of steering hydraulic booster pumps are possible.

NSH32 pumps with belt drive from the crankshaft, completely similar to those used on MAZ cars, are installed on YMZ engines.

NSH32 pumps with belt drive from the engine crankshaft are installed on RENAULT engines. The installation of the pump differs from the suction nozzle with a discharge pipe used on YaMZ engines.

MMZ engines can be equipped with either LH NSH32 pumps or 30A25x136U1 pumps manufactured by Hydraulika BM (Bulgaria) with drive from the engine gear block through an unloading clutch.

The MERSEDESBENZ engines on all bus models and the RENAULT engine on the MAZ105 bus are equipped with pumps installed during the assembly of the engine. In this case, the flow and pressure valve is located on the bracket of the bus frame.

All pumps are equipped with flow and pressure limiting valves used on MAZ vehicles.

Depending on the engine, the spatial position of the flow and pressure limiting valves, the position of the suction pipe with the discharge pipe, as well as its shape, may vary.

On engines with a belt drive of the pump, it is installed on a movable support, which can be rotated on the axis of the fixed support fixed on the engine. Closed ball bearings are installed in movable support, on which pulley is installed. Pulley is fixed on support by locking ring. Pump shaft is inserted into splined hole of pulley, pump is fixed on movable support by four nuts. Flow and pressure limiting valve and suction branch pipe with discharge tube are secured on pump.

Information and patent review of materials on the topic

diploma project

Steering is a set of devices that provide for the rotation of controlled wheels of a car when the driver acts on the steering wheel. It consists of steering gear and steering gear (Fig.1.1). An amplifier may be incorporated into the steering mechanism or drive to facilitate rotation of the wheels. The steering mechanism is designed to transfer force from the driver to the steering gear and to increase the torque applied to the steering wheel. It consists of steering wheel 1, shaft 9 and reduction gear 8. The steering drive serves to transfer the force from the steering gear (gearbox 8) to the controlled wheels of the car and to provide the necessary ratio between their rotation angles.

On cars, a mechanical steering drive is usually used, consisting of a system of levers and rods with hinges: a nozzle 7, a longitudinal rod 6, a lever 5 of a rotary trunnion 4, a transverse rod 3 and rotary levers 2. Levers 2 and transverse rod 3 together with bridge beam form steering trapezoid.

The steering design shall provide:

1) ease of control estimated by force on steering wheel. For cars without an amplifier when driving, this force is 50... 100 N, and with amplifiers - 10... 20 H. For trucks, the force on the steering wheel is regulated by GOST 21398 - 75 and when switching from rectilinear to circumferential movement with a radius of 12 m at a speed of 10 km/h in a horizontal section with a dry hard coating should not exceed: 250 N - for steering control without an amplifier on the way no more than 17 m; 120 N - for steering with an amplifier on the way no more than Nm; 500 N - in case of termination of the amplifier of the path point not more than 17 m;

2) rolling of controlled wheels with minimum lateral withdrawal and sliding at turning of car. Failure to comply with this requirement leads to an acceleration of tire wear and a decrease in the stability of the car when driving;

3) stabilizing the turned steered wheels, ensuring their return to the position corresponding to the rectilinear movement, with the steering wheel released;

4) preventing the transmission of impacts on the steering wheel when the steered wheels hit obstacles;

5) minimum clearances in connections. The angle of free rotation of the steering wheel of a car standing on a dry, hard and even surface in a position corresponding to rectilinear movement is estimated. According to GOST 21398 - 75, this gap should not exceed 15 ° with an amplifier and 5 ° without power steering;

6) absence of self-oscillations of controlled wheels during car operation under any conditions and at any driving modes.

Steering gears

With the help of steering mechanisms, the torque created by the driver on the steering wheel increases. This increase is proportional to the kinematic gear ratio of the steering mechanism Um equal to the quotient from dividing the arbitrarily small rotation angle of the input shaft dα by the corresponding rotation angle of the nozzle shaft deta:

UM=dα/dφ.

The gear ratio can be constant or variable and depends on the design of the steering gear. Although steering gears with variable gear ratio are no more complicated than with constant, they are less technological and therefore more expensive to manufacture.

For steering controls with an amplifier, there are no special requirements on the nature of the gear ratio change, and it is usually constantly. In the absence of an amplifier, it is considered that at the angles of rotation of the steering wheel from its neutral position within ± 90 °, the gear ratio should be increased in order to increase the accuracy of driving and reduce the force on the steering wheel. This is due to the fact that the rotation angles of the steering wheel are mainly in this range. At other steering angles, it is desirable to reduce the gear ratio in order to provide an increased turning speed of the vehicle. Average gear ratios of steering gears of passenger cars Um = 16... 19, freight - Um = 20.5... 23.5.

The steering mechanism shall:

1) be reversible so as not to impede stabilization of the steered wheels;

2) have high efficiency to facilitate control. At the same time, it is advisable to have an efficiency in the forward direction (from the steering shaft to the nozzle shaft) of at least 70% at a moment on the nozzle shaft of 20% of the maximum, and a slightly lower efficiency in the reverse direction.

3) have a minimum number of adjustments;

4) provide a minimum clearance in the middle position of the nozzle shaft corresponding to the rectilinear movement of the car. In this position, the working surfaces of the steering gear parts are subject to the most intense wear, that is, the steering wheel play in the middle position increases faster than in the extreme ones. To prevent jamming in extreme positions during clearance adjustment, steering gear engagement is performed with increased clearance in extreme positions, which is achieved by structural and technological measures. During operation, the difference in engagement clearances in the middle and extreme positions is reduced;

5) ensure the necessary character of the gear ratio change.

The following steering mechanisms are distinguished by the type of gear: worm - roller, worm - gear, worm - sector, screw - crank, screw - rack - gear, gear - rack. In turn, they have a number of modifications.

Worm steering mechanisms are one of the most common and have the following types of engagement: cylindrical worm - central sector, cylindrical worm - lateral sector, globoid worm - two-rib roller, globoid worm - three-rib roller.

The steering mechanisms of the cylindrical worm - the central sector have relatively large dimensions and mass, are characterized by complexity of adjustment, and low efficiency. Variable gap in engagement is provided due to different radii of worm generatrix and initial circumference of sector. In single-axle tractors, this steering mechanism (Fig.1.2) serves only to control the operation of the distributor and therefore the engagement is little loaded. The consequence is that the steering parts are slightly worn out and there is no need to adjust the engagement. Steering gear ratio is constant:

of which is the steering gear sector.

In steering mechanisms, the cylindrical worm is a side sector with relatively small dimensions and weight, the engagement is also little loaded (Fig.1.3). They were installed on heavy vehicles KrAZ and Ural. Their main disadvantage is low efficiency.

Teeth of sector are made spiral. The contact line of the teeth in engagement during rotation of the worm moves along the entire length of the tooth of the sector, which leads to their wear less compared to the other arrangement of the teeth. The engagement is sensitive to changing the distance between the axes of the worm and the sector, so the main parts are made with increased accuracy. Adjustment of engagement is reduced to selection of required thickness of bronze washer installed between support surface of sector and side cover of crankcase. During adjustment, the nozzle shaft together with the sector moves in axial direction. Axial clearance in engagement at rotation of nozzle shaft from middle position to extreme position varies from 0.03 to 0.5 mm.

The worm is cut with a constant pitch t, the gear ratio of this steering gear:

Steering mechanisms globoid worm - roller are widely used on cars and trucks, as well as buses. Compared to other worm mechanisms, this mechanism is more compact and has a higher efficiency, since the contacting engagement surfaces roll around each other, and not slide, as in the steering mechanisms discussed above. These mechanisms are performed with a two- or three-rib roller. The first type (Fig.1.4) is used for installation on cars, and the second - mainly on trucks and buses.

The worm is 5 pressed on a shaft installed in two radial-thrust bearings, adjustable by gaskets 6. To provide adjustment of engagement, the axis of the 8 roller 7 is shifted by 5... 7 mm in relation to the vertical axis of the worm. The engagement is adjusted by a screw 4 screwed from outside the steering mechanism into the upper cover of the crankcase 3. Screw head with adjusting washer 2 enters slot of nozzle shaft head 1. Axial clearance between screw head and shaft slot must be not more than 0.05 mm.

Variable engagement clearance is provided by shifting the axis of rotation O2 of the nozzle shaft by x = 2.5... 5 mm relative to the center of the initial generatrix of the worm O1:

∆3=2∆Rtgδ

To improve engagement, roller rotation axis is turned by angle up to 7 ° relative to worm rotation axis. The lifting angle of the screw line of the worm -α1 is made decreasing from the center to the periphery (ο1min = 5... 10 °), which leads to a decrease in the gear ratio by 5... 7% in the extreme positions of the nozzle shaft. For its middle position:

Uм=2πRw/(tZ1),

where RW is the radius of the worm's initial circle; Z1 - number of worm entries.

Screw steering mechanisms differ in the way of converting the translation of the nut into the angular movement of the nozzle. There are the following types of screw mechanisms: screw - nut - lever, swinging screw - nut, screw - turning nut, screw - nut - rack - sector. Such mechanisms are installed on both cars and trucks. The screw gear is based on the screw driven by the steering wheel shaft and the nut moving along the screw during its rotation. In all types of helical steering mechanisms, the clearance in the helical gear is not adjusted and the helical pair is replaced if the wear limit is reached. Steering mechanisms screw - nut - lever (Fig.1.5) have variable gear ratio, which increases when the nut moves from the middle position to the extreme position:

Uм=2πR/ (tcos2φ),

where R is the distance between the axes of rotation of the screw and lever; t - screw step;

φ - an angle of rotation of the lever (bipod).

The gap in engagement is constant, the efficiency of the steering mechanism is low due to sliding friction between the screw and nut and in connection of the nut with the lever.

Steering mechanism swinging screw - nut (Fig.1.6) has one upper support. When turning the steering wheel, the nut together with the crank moves along the radius R. The gear ratio decreases when the nut moves from the middle position to the extreme position.

and depends on the steering wheel turning to the right (+) or left (-):

Uм=(2πR/t)(sin(γ±φ)/sinγ±cosγsinφ)

where γ is the lifting angle of the screw line of the nut.

This steering mechanism differs little from the previous one, except for the steering column, the dimensions of which are slightly larger due to the fact that the steering shaft swings. There is no adjustment of the axial clearance of the steering shaft. Such a steering mechanism is currently not used.

The steering mechanism screw - rotating nut (Fig.1.7) has small dimensions and weight and can be installed on cars. Clearance between screw and nut is not adjusted. Low efficiency is not essential in terms of steering force when the mechanism is mounted on small cars. This mechanism, like the one discussed above, has a variable gear ratio:

Uм=(2πRcosφ/t±(nRsin(φ)/(n2±(Rcosφk)2).

The steering mechanism of the screw - nut (ball) - rack - is more perfect, which has been widely used in trucks. In the screw pair of these steering mechanisms there is not sliding friction, but rolling friction - between the screw 6 and the nut 5 are placed 90... 120 balls with diameter 7... 9 mm (Fig.1.8, a). Inlet and outlet ends of nut cutting are closed by two guide tubes 4 filled with balls. The result is two closed "streams" in which balls circulate when the screw rotates. Translational movement of nut is converted into angular movement of nozzle by means of rack structurally combined with nut, and shaft of sector (1) with taper fixed on it.

High operability and low wear during the entire service life of such a steering mechanism are achieved by structural and technological measures: high manufacturing accuracy, hardness and cleanliness of working surfaces, the use of highly alloyed steels. The corresponding profile of the mating parts provides a contact spot of the balls on the arc of at least 40... 50 °. The dimensions of the balls differ in diameter by no more than 2 μm, and the gap in the gear does not exceed 0.02... 0.03 mm, which is achieved by selective assembly of parts. In order to obtain a non-locking connection in the middle position of the nut and reduce backshocks, the depth of the screw cutting can be made increasing from the middle to the periphery. For example, in the steering gear of an auto-

mobile MAZ500 on the middle section of the screw 40 mm long, the cutting depth is 0.025 mm less than in the extreme sections. The cutting depth gradually increases in both directions over a length of 55 mm.

The screw-nut connection is not adjustable, which is quite permissible due to the slight wear of the mating parts during operation. The possibility of adjusting the engagement of the rack - sector is ensured by cutting the teeth of the sector at an angle 6... 80 to the axis of rotation, which causes different thickness of the tooth along its length. In addition, the axis of cutting the teeth of the sector is shifted by 0.5... 2 mm with respect to the axis of rotation of the shaft of the sector, as a result of which the thickness of the teeth gradually decreases from the middle to the extreme. This achieves the necessary change in the clearance in the mechanism when turning the steering wheel. This engagement is controlled by rotation of screw 2 whose spherical head rests against washer 3. The nozzle shaft 1 is displaced to the left and the engagement gap is reduced.

In some designs, adjustment of the rack-sector engagement is carried out by approaching the nozzle shaft to the rack using bushings 7, in which the shaft bearings are located (Fig.1.8, b). For this purpose, the center of the circle on which the bolts 8 for attaching the caps 9 to the steering case are located is shifted relative to the axis of rotation of the nozzle shaft. To change the distance between the rack and sector, remove the attachment bolts, turn the covers together with the bushings until their holes coincide with the threaded holes in the crankcase and turn the bolts again. In this case, it is not necessary to cut the sector teeth at an angle to the axis of rotation of the shaft.

Gear ratio of the mechanism is constant:

Uм=2πRw/(t),

where Rw is the radius of the initial circumference of the sector teeth.

In crank steering mechanisms, force is transmitted by means of cylindrical worm 2 and crank 3 with one or two sliding or rotating pins 1 located in screw groove of worm (Fig.1.9). In this mechanism, the gear ratio can be variable and depends on the cutting of the worm. This is an advantage of this type of steering gear, but due to their inherent shortcomings (complexity of manufacture, lack of durability, low efficiency) they are rarely used on modern cars.

Geared steering gears are a conical or cylindrical reduction gear of the planetary or non-planetary type and are sometimes installed on special-purpose machines. They are compact, durable, have high efficiency and constant gear ratio. Gear gears can be used in combination with other types of gears to change the direction of force transfer from the steering wheel to the controlled wheels and to convert the translation of the nut into angular movement of the nozzle. In the first case, a pair of bevel gears (KamAZ, MAZ7310) is used, in the second case, the rake-sector engagement discussed above is used.

In the steering gear - rack (Fig.1.10), the force to the wheels is transmitted using a spur gear or a helical gear installed in bearings and a rack moving in guide bushings. To provide non-locking engagement, the rack is pressed to the gear by springs. The steering gear is connected by a shaft to the steering wheel, and the rack - to two transverse rods, which can be attached in the middle or at the ends of the rack. These mechanisms have a lower gear ratio than the steering mechanisms described above, which makes it possible to quickly turn the steering wheels to the desired position. This led to their widespread use in sports and racing cars, as well as on micro and small cars. Full rotation of the steering wheels from one extreme position to another is carried out behind the 1.75... 2.5 turn of the steering wheel.

The average values ​ ​ of the direct (numerator) and reverse (denominator) efficiency of steering mechanisms of various types are equal: globoid worm - roller - 0.8/0.7; cylindrical worm - sector - 0.7/0.55; worm - crank with rotating finger - 0.75/0.65; worm - crank with sliding finger - 0.55/0.32; screw - nut (sliding) - 0.65/0.25; screw (with ball nut) - rack - gear - 0.87/0.82.

Ferrous metals and non-ferrous alloys are used for the manufacture of steering mechanisms. Steering shafts are made of steels 10, 20, 35, 40, 45. The workpiece is seamless pipes and round steel rolled stock. Globoid worms are forged from steels 35X, 40X, 20KHNM with subsequent cyanidation to a depth of 0.25... 0.5 mm and hardening up to 45... 52 HRC3. Rollers otkovyvat from staly 12HNZA and 20H2N4A, their surface is cemented and heated up to 52...56 HRC3. Worms and rollers are also made of alloyed lead-containing steels ASZOKhM and AS40XhNM, which have increased machinability by cutting. Screws, nuts, laths, waves of a bipod otkovyvat from staly 25HGT, 12H2N4A, 20H2N4A, 20HNZA, ZOH, 35H, ZOHM. Their working surfaces are cemented and crippled to 56... 62 Ue. For production of balls are used steel 25HGT, 12H2N4A, ShH9. As materials for cases and covers serve malleable KCh 3510, KCh 3712 cast iron, steel casting of 08 rcs, 40 l, 45 l and color AK9, AL4, AL9V alloys.

Amplifiers

Amplifiers are designed to reduce the force on the steering wheel during its rotation and to increase "the safety of the car, since the cylinder of the amplifier helps the driver to hold the controlled wheels in a given position when unbalanced forces are acting on the road side, striving to turn these wheels in one direction. Such forces appear when the rolling resistance of the right and left wheels differs due to their hit on the surface of the road with different rolling resistance coefficients (for example, one wheel moves along the sand, the second on asphalt) or in the case of puncture of one wheel.

The amplifier design shall meet a number of requirements:

1) have follow-up action. Distinguish kinematic and power tracking. Kinematic tracking consists in turning of controlled wheels in accordance with steering wheel rotation and its direction. Power tracking ensures that the force on the steering wheel is proportional to the force required to turn the steered wheels, which contributes to more confident driving, especially on slippery roads;

2) provide the possibility of driving in case of failure, amplifier failure;

3) prevent actuation of the amplifier from accidental impacts on the side of the road during straight-line movement of the car;

4) have a high sensitivity, which is estimated by the steering angle corresponding to the increase in pressure in the system to the maximum;

5) have sufficient dynamic stability margin, which is expressed in the absence of self-oscillations of controlled wheels.

Depending on the type of power source used, steering amplifiers are hydraulic and pneumatic, they consist of a power supply unit, a distributor, an actuator and connecting pipelines and hoses. Power supply unit includes hydraulic pump with tank for hydraulic booster or compressor with receivers for pneumatic booster. Some hydraulic booster designs additionally have energy accumulators. With the help of the distributor, energy is supplied to the actuator, which is one or more power cylinders. In them, the energy of compressed air or liquid is converted into a force on the rod transmitted to the controlled wheels of the car.

Currently, only hydraulic amplifiers are used. Since the operating principle of hydraulic and pneumatic amplifiers is the same, only hydraulic amplifiers are discussed below.

The power steering scheme is shown in Fig. By turning the steering wheel 13, for example to the right, the nozzle 12 of the steering gear 14 will turn clockwise and shift the slide valve 9 of the distributor 8 backwards with respect to the received direction of motion of the vehicle. As a result, the liquid from the pump 2 is supplied through the distributor to the cavity A and the power cylinder 7 starts to turn the controlled wheels 4 to the right. Cavity B is also connected through distributor to drain line 1 at this time.

After turning of steering wheel is stopped, controlled wheels continue to turn to the right due to pressure of working fluid on cylinder piston. At that, with the help of lever (5) and rod (3) the distributor housing moves backwards and prevents fluid supply to cavity (A) of booster cylinder, as a result of which rotation of controlled wheels stops. Thus, the steerable wheels are rotated in accordance with the steering of the steering wheel. Kinematic servo action of amplifier is imparted by feedback (lever 5 and rod 3), which connects controlled wheels to distributor housing.

Power tracking action is achieved by introduction of reactive elements: chambers or plungers. Most amplifiers mounted on modern cars have not only kinematic, but also power tracking. In Fig.1.11 power tracking is achieved by means of jet chambers 6 and 10, into which liquid is supplied through calibrated holes from pressure line and acts on right or left end of slide valve 9 depending on direction of vehicle rotation. As a result, the force required to move the spool is dependent on the pressure in the pressure line 11, which in turn is determined by the moment of resistance to rotation of the controlled wheels. With its increase, the pressure in the cylinder and in the jet chamber of the distributor increases, preventing the slide valve from shifting and facilitating its installation in a neutral position.

Inertia of the steered wheels at unsuccessful amplifier parameters can cause further displacement of the distributor housing relative to

1 - distributor; 2 - power cylinder; 3 - steering mechanism.

slide valve (transition through neutral position). In this case, the pressure line will be connected to the cylinder cavity B and the wheels will turn in the reverse direction, that is, under certain conditions, self-oscillations of the controlled wheels can occur. The presence of reactive elements in the distributor reduces the likelihood of such oscillations.

Depending on the relative arrangement of the elements, four amplifier layout diagrams are distinguished (Fig.1.12). In all these circuits, the power source (pump) is located separately from the other elements of the amplifier.

When the distributor and cylinder are located in the same unit as the steering gear (Fig.1.12, a), the design is called a hydraulic steering wheel. Its advantages lie in compactness, a minimum number of hoses and pipelines, a low tendency of the system to self-oscillation due to the high rigidity of hydraulic lines connecting the distributor with the power cylinder. However, in such a structure, the entire steering drive from the nozzle to the steered wheels is loaded by an additional force applied from the cylinder side to the nozzle shaft. This leads to the need to increase the size and weight of the actuator. The hydraulic rudder has large overall dimensions, which makes it difficult to arrange it on a car.

In addition, hydraulic rudders are inconvenient from the point of view of unification of steering elements. Nevertheless, hydraulic rudders are widespread on cars and trucks, as well as on buses.

The arrangement of the amplifier (Fig.1.12, b) is characterized by the placement of the distributor in one unit with the steering mechanism and the autonomous arrangement of the cylinder. This allows the cylinder to be mounted in close proximity to the steerable wheels. The advantages of an amplifier with this circuit lie in the low load of the drive, the ease of arrangement of the amplifier in the steering drive, and the low tendency to self-oscillations. The cylinder located near the wheels perceives impacts from the side of the road, protecting the steering mechanism from overloads. When using this amplifier circuit, the length of the hoses is slightly increased compared to the previous one.

According to the diagram given in Figure 1.12, the steering mechanism is installed autonomously, and the distributor and the power cylinder are installed together. In this case, the cylinder must be arranged in strict accordance with the steering mechanism, since the ball pin of the nozzle must control the operation of the distributor. Amplifiers made according to this scheme have a low tendency to self-oscillations. The length of pipelines is slightly increased compared to the previous design.

The steering scheme with the independent arrangement of the steering mechanism, distributor and power cylinder (fig. 1.12, d) is the most flexible in terms of the layout and unification of the elements. However, due to the increased tendency to self-oscillations, the increased number and length of hoses and pipelines, it is relatively rare to use.

The layout of the amplifier and its characteristics are chosen mainly depending on the load on the controlled wheels. In addition, constructive and technological continuity is taken into account.

The performance of the hydraulic booster depends on the design of the distributor. Distributors are:

1) of open and closed types, in the first case the width of the spool edges is less than the width of the corresponding holes in the housing. As a result, during straight-line movement of the car, the pressure and drain lines of the amplifier through the distributor are connected to the working cavities of the cylinder. Since the booster pump is constantly operating, the liquid is continuously circulated through the distributor. In distributors of the second type in neutral position of slide valve all lines are closed. Fluid is supplied to distributor-divider from hydraulic accumulator. The amplifier pump is switched on periodically and serves to recharge the hydraulic accumulator. Such a system allows the use of a pump with a smaller supply and reduces the energy consumption of its drive;

2) with axial or angular movement of the slide valve. Currently, distributors of the first type are more common. Distributors with angular movement of the slide valve are characterized by high sensitivity and ease of drive;

3) with and without reactive elements;

4) with self-aligning slide valve or with its centering by means of elastic elements (springs, torsion bar). In the first case, the centering is carried out due to the action of the liquid on the reactive elements, in the second - when the slide valve is shifted, a force occurs from the side of the elastic elements, which tends to return the slide valve to a neutral position.

Distributors of three types became most widespread: with reactive plungers and centering pre-compressed springs (type A); with jet chambers and self-aligning slide valve (type B); without reactive elements with centering pre-compressed springs (type B). The simplest is type B distributor, which provides only kinematic tracking (Fig. 1.13). The slide valve 2 is moved relative to the sleeve 1 by means of a ball pin 4 connected to the bevel of the steering mechanism. For this purpose it is necessary to overcome the force of spring 3 installed with preliminary tension. At that pressure line H is connected to one of cavities of power cylinder, and drain line C - to the other. When the steering wheel is released, the spring returns the slide valve to the middle position, in which the pressure and drain lines are connected to each other and to the working cavities of the cylinder. In the event of a failure of the amplifier, the ability to control the car with a significant increase in force on the steering wheel remains. To reduce hydraulic losses in case of non-operating pump, ball valve is built into distributor housing, through which liquid flows from one cylinder cavity to another when car turns.

The Type B distributor design is shown in Figure 1.14. Pins (5) and (4) are attached respectively to ploughshare and longitudinal tie-rod, and housing (3) - to cylinder housing, rod of which is fixed on LH spar of frame. The nipple pin 5 can displace the spool 1 axially by means of the sleeve 2 and the tie rod 6 by the amount of clearance S.

Jet areas of slide valve are made proportional to working areas of power cylinder to the right and to the left of piston. As a result, the slide valve is slightly shifted from the middle position towards the reaction chamber with a smaller area during straight-line movement of the car. This makes the forces acting on the piston on the right and left equal at different fluid pressures. This distributor operates in accordance with the diagram shown in Fig.1.11.

Due to the absence of centering elastic elements, the force required to move the slide valve relative to the middle position is insignificant. As a result, B-type distributors are highly sensitive to accidental road effects, which can lead to self-actuation of the amplifier and the influence of controlled wheels. This drawback has been eliminated in Type A distributors with reactive plungers and centering pre-compressed springs. Despite the more complex design, they are widely used on cars and are preferred in the development of amplifiers for promising cars. These structures, like those discussed above, refer to axial displacement distributors.

To displace the slide valve 6, it is necessary to additionally move the plungers 3 to the left, which are influenced by the forces of the springs 2 and the pressure of the liquid, which tends to return the slide valve to the middle position. Otherwise, the operation of this amplifier is similar to that of the amplifiers discussed above.

promising MAZ cars.

In the distributor located in the same housing with the steering gear, the screw 5 installed in the radial thrust roller bearings does not have axial movement (Fig. 1.16). When the steering wheel is turned, the screw bushing 3 with the spool 2 fixed on it moves in the axial direction, as a result of which the amplifier begins to work. Displacement of slide valve is determined by angular clearance in splined joint of shaft 4. To hold slide valve in middle position there is torsion bar 1 connecting shaft and screw. Otherwise, the operation of this distributor is not fundamentally different from those discussed above.

Type A distributors ensure equal pressure in cylinder working cavities. Since the active area of the piston on the side of the rod is smaller, during straight movement a small force is created that tends to turn the controlled wheels in one direction.

The torque on the steering wheel from the action of the distributor reactive elements in passenger cars can reach 1/3 of the moment of resistance to turning the wheels. In order to maintain the servo action and to avoid excessive force on the steering wheel, some amplifiers introduce devices limiting the reactive action of the distributor.

All distributors discussed above have axial movement of slide valve. In the most common amplifier circuits (Fig. 1.12, a, b), the slide valve drive is quite complex, which is a significant drawback of the design. The drive is significantly simplified if you use a spool not with axial, but with angular movement (rotary). Such amplifiers have little hydraulic play, since in this case the slide valve is directly connected to the lower end of the steering shaft. It is difficult to impart force tracking to the rotor distributor, which is a disadvantage thereof.

Fig.1.17 shows an embodiment of the rotary distributor. Slide valve 2 has four long longitudinal slots connected to pressure line and four short slots connected to drain by radial holes. When the steering wheel is released, the slide valve is held in the middle position by a torsion bar 1 connecting the slide valve and the steering screw 4. Screw has lock-free connection with sleeve 3 by means of pin 5 and connection with slide valve by means of end teeth allowing relative angular displacement of slide valve and sleeve up to 3 °. In this way, the slide valve is displaced relative to the sleeve to actuate the amplifier.

When turning the steering wheel, the spool rotates relative to the sleeve by an angle of 3 °, twisting the torsion bar and including the amplifier. At further rotation of steering wheel together with slide valve sleeve rotates relative to housing 6. The operation of the distributor is understood from the figure, it does not fundamentally differ from the operation of the distributors discussed above with axial movement of the slide valve.

Steering hydraulic booster pumps shall provide the necessary feed as determined by the steering wheel design speed; have a small change in supply at different engine speed modes; provide the required pressure; have sufficient durability and reliability.

Pumps operate under unfavourable conditions with drastically varying speed and load modes. The pump is driven from the engine, and the nature of the change in their speed modes is the same. Pressure developed by pump is proportional to torque on controlled wheels. The change in pressure is random and depends not only on the rotation of the steered wheels, but also on their blows about the irregularity of the road during rectilinear movement of the car. Operating fluid temperature in hydraulic pump tank can range from + 100 ° С in summer to - 50 ° С in winter. The medium in which the pump operates is characterized by increased smoke and dust content. These operating conditions determine the design and type of pumps used.

Controlled bridge beams

These beams shall have the necessary strength and rigidity. The beams of the continuous front bridges have undergone relatively small structural changes. When the production of cars was small I-beams of front bridges were cast from steel or ductile cast iron. On some models of cars, tubular continuous front bridges made of chromomolybdenum steel were installed. However, the latter in mass production had a higher cost than the previous ones.

On cars of the first release, only the rear wheels were equipped with brakes. The absence of front wheel brakes significantly reduced the loading of the front axle beam with a bending moment in the horizontal plane. Note here that front position beam is loaded mainly by bending moment in vertical plane and, to a much lesser extent, horizontal one.

The traditional manufacturing of front bridge beams by forging is due to their complex configuration, which makes it difficult to use another method, and also to the fact that the fatigue resistance of such beams is high, which is due to the formation of a fiber-oriented structure along the axis of the beam.

With the use of front wheel brakes and an increase in their efficiency, the proportion of the horizontal bending plane and torques perceived by the beams of the front bridges increased. Therefore, an I-section for these beams should not be considered optimal, since it is advisable to use it for beams that perceive bending moment mainly in one plane only - vertical.

The vertical wall of the I-beam is practically not loaded when bending in the horizontal plane, and all loads are perceived only by horizontal shelves, the cross-sectional areas of which are small and slightly distant from the neutral line. Considering the ratio of bending moments perceived by the beam in the vertical and horizontal planes under real conditions of movement, it should be recognized that the most rational cross-section for the front axle beam is a rectangular hollow section with a larger side in the vertical plane or an even more efficient differential rectangular hollow section with narrow horizontal walls of large thickness. The latter section provides optimal ratio of moments of section resistance to bends in vertical and horizontal planes and sufficiently well withstands torsional loads. However, such sections are currently used only for rectilinear beams (for bridges of trailed links of road trains). As a result of the fact that the beams of the front bridges of cars have bends, they are made with an I-beam cross section. The exception is welded beams of front bridges of heavy vehicles, in which the middle part of the beam has a straight hollow tubular section, and forged ends - I-beam.

The I-section of the front axle beams differs significantly from the standardized I-section. Modern forged steel beams of front bridges are a solid part (Fig.1.18). They have platforms 2 for attachment of springs and bosses 1, in which holes are made for installation of pivots.

The parts of the beam between the spring platforms and the pivot holes are conventionally called external, and the part of the beam between the spring platforms is conventionally called middle. From the layout conditions and to ensure the necessary working clearances between the engine and the beam (especially in the case of suspension breaks in case of compression limiters failure), the latter has bends in the outer parts (GAZ 53 car), and sometimes both in the outer and middle parts (MAZ503 car).

The shapes of the I-sections of the beams of the front bridges are diverse. The most common cross sections are shown in Fig.1.18. Horizontal surfaces of upper and lower flanges of section have convex shape with rounded edges. The slope of the shelf must be at least 7 ° so that the forging is easily separated from the die. The dimensions of the horizontal shelves and vertical walls, as well as the radii of the rounding, differ significantly from the dimensions of the standardized I-beams and for different beams of the front bridges, close in load capacity, have different values.

The cross-sectional dimensions of the front axle beams along the length are usually not constant (see Fig. 1.19). Thickness of vertical section wall gradually increases as it approaches pivot hole to form pivot boss 1. Upper horizontal flanges of outer and middle parts of beams are smoothly narrowed from spring platforms to pivot bosses. Spring platforms at the point of transition to the upper shelves of the I-section of the middle part of the beam have relatively large radii of curvature.

Radii of curvature of sections of transition from spring platforms to horizontal flanges of outer and middle parts for different beams are different (see Fig. 1.19). The beams of cars intended for use on roads with improved coating make the radii of curvature smaller than the beams of cars intended for use on bad roads. External parts of beams may have local bulges and holes for attachment of shock absorbers. Lugs are made near pivot bosses to stop bolts limiting maximum rotation angles of wheels. The cross-sectional dimensions of the middle part of the beam (between the spring pads) are usually constant.

During forging, it is necessary to provide rounded I-section shapes that increase the durability of the dies and facilitate separation of the blank from the die surface.

The characteristic features of the cross sections of the beams of the front bridges are due to the desire to reduce the redistribution of metal during forging and, as a result, increase the resistance of the dies. The latter, in particular, explains the widespread use of beams with cross section II (Fig. 1.18). Such beams are installed on FordTransit 900 cars, etc. Smooth transitions in the section of the beams of the front bridges cause a decrease in stress concentrations. Beams with section I have better strength characteristics than beams with section III, since in such section a large mass of metal is concentrated on peripheral sections of horizontal shelves, the most distant from neutral lines during bending in vertical and horizontal planes. However, in a beam with section III, the metal is redistributed less, the stamping slopes are larger, which increases the durability of the dies.

Some I-beams are made by partial rolling. The production process begins with rolling of a block of rectangular section to obtain an I-beam. The beam is then shaped and the ends are planted and stamped in closed dies. Rolling slightly increases the strength qualities of the front axle beams.

Usually beams are made by forging. Forging is first subjected to heat treatment to relieve internal stresses, and then to cutting treatment.

Some dimensions of cross sections of beams of front bridges are regulated by OST. The standard does not apply to split beams of front bridges of cars intended for off-road operation, vehicles of small-scale production, vehicles whose maximum speed does not exceed 25 km/h and which have less than four wheels if their total mass does not exceed 1 t.

Pivot holes in beams are of two types: cylindrical

(see Fig.1.19, a, b) and conical (see Fig.1.19, c). The former provide reduction of manufacturing cost, and the latter make it possible to increase reliability of pivots attachment in beams.

The surfaces for the nuts of the ladders are usually processed. Sometimes these surfaces are produced without slopes and are not subjected to cutting treatment, which reduces the cost of making beams. An important characteristic of the front axle beam is its bending rigidity in the vertical plane. At present, sufficient reasonable requirements for the rigidity of the front axle beams have not yet been developed and the necessary statistical material has not been accumulated. When assessing the stiffness "of a beam of a designed car, it should be compared with the stiffness of beams of similar cars, the reliability of which has been checked by experience in operation. Therefore, when choosing the dimensions of the front axle beam and its material, the designer must find the optimal solution taking into account the operating conditions of the designed car and providing the necessary stiffness of the beam at the minimum weight.

Of the published statistical materials characterizing the strength of the front axle beam, the dependence given in Fig.1.20. is of interest. When determining the moments of resistance to bending, the I-section shelves having stamping slopes were replaced by rectangles equal in area and length. At the same time, the positions of the centers of gravity adopted during the calculation of the shelves relative to the center of gravity of the entire section were preserved. The constraint shown in Fig. 1.20 shows the linear relation of the resistance moment, beam section with axial load. Most front bridge beams are made of carbon steel 40, 45, chromium 30X, 35X or chromium nickel.

When using carbon steel, the dimensions of the beam sections and its mass increase, but the necessary rigidity is ensured. The use of chromium and chromium-nickel steels allows to reduce the weight of the beam and increase its durability, but at the same time a slight decrease in its stiffness is possible. So, for example, replacing steel 45 with steel 40X in the manufacture of the front axle beam of the ZIL130 car allows you to save metal 8-9 kg.

The desire to reduce the unsprung masses of the car has led to the recent use of aluminum alloys for beams of front bridges of trucks and truck tractors. Such forged aluminum beams are used on some cars with an axial load of 47.3 to 54.2 kN, for example, the Volvo (Sweden) FordSL900 and International Harvester (USA). A significant reduction in beam weight (by 45 kg) in this case is achieved while maintaining high strength and durability. High-strength heat-treated aluminium alloy is used as material for such beams. The yield strength of this alloy is close to the corresponding values of the yield strength of the steels used for the manufacture of beams. The aluminum beams also have an I-section. For the manufacture of aluminum beams, alloys containing Al, Mg, Si, Cu, Mn are used. The Si content of these alloys sometimes reaches 1% and Mg 3%.

Turning knuckles, pivots and wheel hubs

Turning knuckles, pivots and hubs of the wheels of the front axles of cars are one of the most important parts, for strength and durability, which are given increased requirements. Turning knuckles (Fig.1.21) consist of trunnion (1), flange (2), upper and lower pivot lugs (3) and holes (A) for attachment of side levers of trapezium and steering rod (upper hole is available only for left pivot knuckles and serves for attachment of lever of longitudinal steering rod).

The trunnion starts with the collar 1 from the flange and has two cylindrical surfaces for installation of internal races of bearings, a cone surface at transition from the mounting surface of the internal (larger) bearing to the mounting surface of the external (smaller) bearing and ends with a thread.

The design of the flanges of the turning fists is different: the flanges are round, square and rectangular. Round flanges have, as a rule, six or eight holes, and square and rectangular - four or six. The higher the number of threaded holes for attachment of brake support discs, the higher the reliability of this connection. However, this increases the labor input of the assembly and the cost of manufacturing the rotary knuckles. In some foreign designs, the flanges do not have threaded holes, since the brake support disk is attached to the flange of the rotary knuckle, forming an integral connection. This makes it possible to slightly reduce the labor input of fabrication and assembly.

The upper and lower parts of the flange of the swivel cam pass into pivot eyes with a cylindrical surface located at an angle to the flange plane, which provides the necessary tilts of the pivot. Holes for attachment of side levers of trapezium and upper lever of longitudinal steering rod have conical surface with key grooves. Radii of fillets at the point of transition from flange to pivot eyes provide elimination of zones with abrupt change of stresses.

Swivel knuckles are made by forging from doped chromium-nickel or chromium steels. In this case, the fibrous structure of the swivel cam material is provided, which causes its high fatigue resistance. Forging is processed on automatic lines and then subjected to heat treatment.

The main fatigue breakages of the rotary knuckles (about 70%) occur in the galley area at the transition from the trunnion to the flange. They exceed the endurance limit of this interface using various structural and technological measures. Thus, the experience of the operation of ZIL130 cars and LiAZ, LAZ and Ikarus buses, which use forging of the turning fists of the ZIL130 car, showed the insufficient durability (in the place of the galley) of the latter when installing them on buses, since in this case they carry loads greater than on the ZIL130 car. The use of pebble rolling increased the endurance limit of this mating by 50% and durability by 4.3 times while maintaining the same dimensions of the rotary fist. On MAZ cars, to increase the fatigue resistance of the zone, its radius was almost doubled (up to 12 mm). To increase the fatigue resistance of the rotary fists of these cars, heat treatment of the surface of the TVH trunnion with deep heating was also used. Studies have shown that such a heat treatment allows increasing the fatigue resistance of the rotary knuckles by 20% compared to volume quenching. In this case, the surface hardness of the rotary cam reaches HB 290-320. The turning fists of the MAZ500 car also increased the flange thickness. The fatigue resistance of the turning fists of these cars ensures their mileage up to 400 thousand km.

In some rotary knuckle designs (e.g. PAZ672 bus), the fatigue resistance stress of the pivot assembly was observed due to insufficient bridge thickness between the cone hole surface and the thrust bearing seat. Fatigue resistance of such swivel knuckles was increased as a result of increasing the wall thickness of the pivot assembly at the point of connection with the steering trapezium lever and eliminating slots from the tool when treating the ends of the eyes. At the same time, it is also advisable to use rolling with a roller of the tapered surface of the turning cam under the lever of the steering trapezium.

Pivots, as a rule, are cylindrical in shape and are made of various steels. The advantages of cylindrical pivots are the simplicity of manufacture, the high reliability of attachment due to the elimination of the possibility of their rotation in the beam and the small number of attachment parts.

Some heavy-duty trucks use stepped pivots with a threaded tail (Figure 1.22). Such pivots are installed in cone sockets of beam by means of nut and spacer bushing, which rests against upper end of beam.

In most cars, pivots are pressed into the pivot holes of the beams and locked in them with one or two wedge pins. For this purpose, the pivots have two or four flats, which allow during operation to turn them relative to the turning cam, since the working surfaces of the pivots wear out unevenly. When the pivot is rotated, the required clearances in this mating are restored without replacing the parts.

In some cars (for example, MagirusDeitz), additional wedge locking of pivots was not carried out, and during the operation of cars in heavy road conditions, cases of pivots falling out were observed due to significant elastic deformations of the beams. In this regard, it became necessary to introduce a protruding belt into the structure of the pivot, which significantly complicated its manufacture.

The working surfaces of the pivots are cemented to a depth of 11.5 mm before grinding to a hardness of HRC 56-63 or subjected to surface hardening of the HPV to a depth of 2-2.5 mm. Such heat treatment significantly increases the wear resistance of the pivot without reducing its fatigue resistance as a result of the formation of a viscous core in the pivot material.

Currently, chrome plating of pivots is used, which increases their wear resistance and increases service life. It should be noted that the pivot coupling (like the tips of the steering rods) in the front bridge wears out most quickly. This should be taken into account especially when designing cars designed for operation on bad roads.

When making pivots from chromium-nickel steel, surfaces that are in contact with antifriction bushings pressed into eyes of turning knuckles are significantly strengthened. Steel tape coated with antifriction layer of lead or tin bronze is used for bushings. In some cases, phosphorous bronze is used. To increase the service life of the pivot coupling, needle bearings are used, as well as pivot bushings made of powder materials that well retain the lubricant and allow increasing the frequency of lubrication of the coupling.

Wheel hubs are divided into two main types: spoke (or flange - for disc wheels) and spoke (for disc-free wheels). Flanged hubs are used on cars and light and medium load trucks. Flange hubs 4 (Fig.1.22) are of complex shape and consist of bearing part 4 (I) (with bearings 2) and flange part 4 (II). Flange part has holes into which wheels attachment studs I are pressed. On the inner side of the flange part of the passenger car hubs, the belt A is leaked, with which the flats of the wheel attachment studs heads are in contact. As a result, the studs are prevented from turning when the wheel nuts are tightened. For trucks (Fig.1.23, b), studs I are tightened by nuts 8 on the inner side of the flange part of the hub. Flange part of hubs of such vehicles has thickening at locations of pins and stiffening ribs 7 increasing rigidity of connection of flange part with bearing part. The following structural and technological measures are used to reduce the loaded fillets and stiffeners at the point of transition of the bearing part to the flange part, as well as to increase the strength of the hubs: increase the thickness of the flange part in the absence of stiffeners; eliminating the undercut of the flange portion during cutting; increase the radius of the spindle (up to 3-5 mm), the thickness of the stiffening ribs, the radii of transition from the surface of the ribs to the bearing and flange parts of the hub (to reduce stress concentration); quality of hub blanks casting is controlled to eliminate shells and loosening and quality of their cutting treatment.

On the outer side of the bearing part of the hubs there is provided the attachment of the cap 5, which serves to retain the lubricant in the bearings 2 and to protect this unit from dirt ingress. Gland 3 prevents hot lubricant from getting from bearing assembly on brake parts. Nuts (6) for bearing attachment in passenger cars are locked by cotter pin, and in trucks - by lock washer with lock nut or threaded clamp.

For wheel hubs, mainly conical roller bearings are currently used. The increase in the speed of cars and road trains on motorways led to a revision of the operating conditions of wheel hub bearings. Some foreign firms have developed the designs of such seals,

which allow the use of liquid lubricant for these bearings, improve their operating conditions and increase their durability. The bearings operate in an oil bath which is located between the inner bearing seal and the transparent plastic cover. The cover is made without a vent hole, as a result of which moisture and dust are excluded from entering the bearings. Typically, the outer bearing is smaller than the inner bearing in the hubs. In order to improve the reliability of this unit, Frühauf sets the external and internal bearings of the same type on the trailer axle journals designed for an axial load of 111.1 kN. Bearings in this case are designed for a radial load of 196 kN, which is 25% higher than in other bridges of similar lifting capacity.

Naves of wheels of bridges of cars produce from steel or pig-iron (KCh 3712 malleable cast iron) of cast preparations, processed by cutting on automatic transfer lines. Places of pressing-in of external races of bearings and gland, as well as flange part with which brake drum and wheel disc contact, are subjected to cutting treatment. A new trend is the manufacture of passenger car hubs together with a bearing assembly. Hub is used as external collars of bearings. Such a design requires an improvement in the quality of the hub material and the accuracy of its processing, as well as heat treatment to obtain the necessary hardness of the running tracks of the bearing rotation bodies. In this case, when the raceways are worn out, the hubs are replaced together with the bearings. Along with some complexity in the production of such hubs, the overall labour input of bearing assemblies and the mass of the hub are reduced.

For bridges of trailed links of road trains, hubs or specially designed or the same as on bridges of cars are used. Special hubs are only useful for repetitive production. At the same time, on semi-trailers and trailers with single wheels, a special "trailer" hub can be made lighter as a result of the use of a wheel whose disk has a zero and close to zero departure.

The departure of the wheel disk causes a significant moment (from the vertical reaction), bending the disk part of the hub relative to the bearing one even in straight line movement. So, the departure of the wheel disc from the GAZ53A car is 13.2 cm, and from the ZIL130 it reaches 16.8 cm. If the same disks are installed in the cars on the front and rear bridges, the wheels have a large disc departure (8-17 cm), which causes the need to use on bridges with single wheels not quite rational and heavier hub design. Therefore, the use of such hubs on the axles of controlled-wheel vehicles is a compelled compromise solution that is not always appropriate for single wheels of trailed links.

Patent Review

Steering.

Application 10159326 Germany, B 62 D 5/06.

The proposed steering has high functional reliability. It consists of a steering wheel, a shaft with a ball hinge or rack gear control, a rack gear mechanism, a lever and a transverse link of the steering trapezoid, as well as a hydraulic control device. Transverse tie-rod made integral with mechanism rack represents double-sided hydraulic cylinder of actuator whose cavities are connected with control device.

This design optimizes the functions of both systems and ensures compact execution. In this electrohydraulic, electromechanical or pulse open control system built in or part of the brake system, the external energy source of the drive is a hydraulic energy accumulator with control of the action on the actuating hydraulic cylinder of the steering wheels by means of two de-energized closed and two de-energized open two-way proportional electromagnetic valves (see Fig.

Steering system.

Application 10046168 Germany, B 62 D 5/04.

The proposed system of indirect steering of a road vehicle provides its automatic operation. To this end, the system is equipped with hydraulic and electric amplifiers. Torque is transmitted from steering wheel by split shaft of steering column connected by clutch to gear of steering mechanism with toothed rack, which is extension of cylinder-piston hydraulic booster. Oil is supplied by hydraulic pump to its cavities via oil pipelines opened and closed by solenoid valve. In this mode of hydraulic booster, split shaft half-couplings are attracted to each other by magnet. In the electric booster mode, the half-couplings are open and the torque transmission from the steering wheel is monitored by an upper electric motor activator, the gear of which is in engagement with the gear rim of the half-coupling. Rotation of the drive rotor leads to excitation of the induction current in the control winding of the compact unit, which also includes the second lower motor of the drive. Its rotor, rotating under the influence of electromagnetic induction, performs the function of an electric amplifier (see Fig. 1.25).

Rotor type distributor.

6,119,803, B 62 D 5/10.

A hydraulic booster with a rotary spool is patented, which does not create noise at the position of the steering wheel in neutral and provides a better reaction when it turns from neutral. In the slide valve there is a pair of clearances through which fluid is supplied under pressure. Through the other pair of clearances, liquid enters one or another cavity of the hydraulic cylinder of the amplifier. Through a pair of gaps, the source of fluid under pressure is connected to the drain and through a pair of gaps, this or that cavity of the hydraulic cylinder is connected to the drain. The gaps of each pair are 180 ° offset from each other, i.e. they are opposite to each other. At neutral position of steering wheel the value of closing of clearances through which fluid under pressure enters one or another cavity of hydraulic cylinder exceeds the value of clearances through which specified cavities of hydraulic cylinder are connected to drain. This solution allows the driver to feel his reaction better when turning the steering wheel slightly from the neutral position (see Figure 1.26).

Pivot knot.

Application 93033349, B 62 D 7/18.

A pivot assembly, preferably a vehicle wheel rotation mechanism, comprising a pivotal trunnion with lugs spaced along the pivot axis; made in the form of terminal clamps with coaxial holes in them; a bearing part with a hole and a pivot mounted in the holes of the journal tabs and the bearing part; on which on one side there is a threaded element, characterized in that: that in coaxial holes of lugs there is a thread with the possibility of pivots fixation along mating threaded surfaces, on the ends of which there is a polyhedron (see Fig. 1.27).

Installation of pivot pin.

Application 10161207 Germany, B 62 D 7/18.

The invention proposes an improved effective unattended self-regulating structure of a pivot unit for a front driving axle and a non-driving axle connecting them to controlled wheels. In the first case (Fig.1.28, a), pivots inserted and fixed in the sockets of ball support of axle stockings serve as axes of rotation of turning trunnions of wheels. In the second case (Fig.1.28, b) the single pivot fixed relative to the rigid beam is the axis of rotation of the turning cam. At that in through inclined-vertical holes of rotary journal and rotary cam pivots/pivot are installed by means of needle rolling bearings and machine-shaped end sliding bearings. The latter are pressed to ends of pivots through compressed elastic gasket by threaded sealing plugs - plugs with welded hexagonal heads. Grease is placed inside the interface and glands are inserted.

Driven axle of vehicle.

Application 10133937 Germany, B 62 D 7/00, B 60 G 7/00.

The proposed controlled axle differs from traditional designs by reduced relative movement of steering rods and a swivel knuckle. The hot-stamped rigid axle beam with spring or air suspension with similar longitudinal levers docked to it by four screws is connected to the car body. On the other end of one of the levers, located under and parallel to the longitudinal steering rod, the crankcase of the steering mechanism is fixed. All these structural elements are subassemblied into the pre-installation module. When using a conventional steering mechanism, the steering nozzle is pivotally connected to the longitudinal steering rod. Instead, a working actuating cylinder fixed to the end of the longitudinal suspension lever may be used. The end of the longitudinal steering rod or rod of the actuating cylinder of the steering trapezoid is pivotally connected to the bracket of the rotary lever, the other end is fixed on the rotary cam (Fig.1.29).

Justification and design selection of designed units

Based on the requirements set for the bus steering structure, taking into account the information available on this topic, we accept the steering scheme presented in Fig.2.1.

Torque from the steering wheel through the steering wheel shaft, an intermediate angular reduction gear changing only the direction of torque, the intermediate cardan steering shaft is transmitted to an integral steering mechanism with a rotor-type power distributor. In the steering mechanism, through the screw-nut pair, the rotary movement of the screw is converted into the translational movement of the piston, into which the nut is mounted. Force (torque) is transmitted to sector shaft through gear engagement of piston-rack with toothed sector.

Steering mechanism nozzle pivotally connected with longitudinal steering rod is secured on sector shaft. Thus, the force from the steering mechanism through the lug and the longitudinal steering rod is transmitted to the lever of the turning cam of the right wheel of the bus. The right and left wheels are interconnected by means of a steering trapezoid consisting of two (right and left) levers of the steering trapezoid and a transverse link of the steering trapezoid. As noted, the steering to reduce driver fatigue in conditions of intensive urban traffic is equipped with a hydraulic amplifier. The adopted steering scheme provides relatively low saturation of units and parts, which has its own positive aspects during operation and repair.

The design of the beam of the front non-driving axle should provide stability of the installation angles of the controlled wheels and the necessary kinematics of their rotation at all modes of the bus movement. The most responsible from this point of view is the steering assembly of the controlled bridge. The pivoting joints of the pivot pin and the swivel cam shall be designed so that the mounting angles of the wheels to be rotated during operation are not altered or easily restored by appropriate adjustments.

The shape of the beam should be such that the required ground clearance and low arrangement of the units relative to the load-bearing part of the bus are provided.

This condition is important, since the driver's workplace in the bus in question is advanced and lowered downwards. In addition to the requirements discussed above, the bridge beam is subject to strict strength requirements. For this reason, the beams of controlled bridges are traditionally made in the form of forging of an I-section (the manufacture of beams of such a section is most technologically, plus, the required safety margin is provided) as a whole with the knuckles for fixing the pivot and with the platforms for attaching the pneumatic suspension elements. Diagram of the section and shape of the received beam of the front non-leading bridge are shown in Fig. 2.2, a, Fig. 2.2, b, respectively.

Selection of basic parameters and calculation of designed nodes

3.1 Selection of main steering parameters

The steering parameters are selected based on the technical characteristics of the MAZ103 bus.

3.2 Kinematic and power steering calculation

Kinematic calculation of steering consists in determination of rotation angles of controlled wheels, selection of parameters of steering trapezoid and check of trapezoid for jamming. The design diagram is shown in Fig.3.1 and Fig.3.2.

Maximum rotation angle of the inner wheel

Calculation of curve of specified change of wheel rotation angles

Angle δa max is calculated by equation:

where j - distance between pivots of controlled wheels;

as a function of the given value δi max:

Then, using the same equation, intermediate values are determined

δa = f (δi) at iterative increase δi each time by 5 °. The calculation results are summarized in Table 2.

Approximate determination of rotation angle of steering trapezium lever α1

This calculation is made according to the equation:

The positive error, which can hardly be tolerated, should, however, disappear, given the inclined position of the transverse thrust in the rear view.

Consideration of the angle of lateral inclination of pivot axles α0 at the rear view. Determination of angles [omega] and [omega] 3, as well as errors ∆u

The design diagram is shown in Fig.3.3.

The difference between α2 and α3 is. When calculating, all steering and front suspension elements are absolutely rigid, that is, we neglect the available compliance. Therefore, in the drawing, the angle of rotation of the steering trapezoid lever must be indicated with the tolerance:. This ensures that even the maximum rotation angle of the outer wheel is not less.

Check of the largest, and smallest, angles between the transverse tie-rod and the steering trapezium lever

The angles shall be within 15 ° 165 ° so that the lever drive does not jam.

Determination of actual rotation angle of outer wheel δa

This angle can only be found by equations:

The actual rotation angle of the outer wheel is.

Error:.

The results of the calculation of the above equations taking into account the slope of the transverse thrust in the rear view are given in Table 2 for values ​ ​ discretely increasing by 5 °. To illustrate the effect of the tolerance field on, the results of the calculation for its extreme values and

Based on the obtained results, we will build dependencies of outer wheel turn angle from inner wheel (Fig.3.5).

Power calculation of steering consists in determination of moment arising from action of rolling resistance forces, moments from friction of drilling on controlled wheels, moment of lifting resistance, and also in determination of force applied by driver to steering wheel for wheel rotation.

In view of the fact that in modern steering structures the force on the steering wheel should be within 120160 N, it is necessary to use an amplifier in the calculated steering. The best option for the bus is hydraulic power steering.

3.3 Static and Hydraulic Calculations

steering hydraulic booster

We define the work balance:

The diameter of the power cylinder is determined based on the required movement of the longitudinal steering rod, when turning the wheels by the maximum angle, and therefore, taking into account the ratio of the steering rod and the piston of the hydraulic cylinder. The movement of the steering piston-rack can be determined by the expression:

Hydraulic part requirements:

• spool leaks do not exceed 510% of pump capacity;

• slide valve stroke must not take into account steering wheel play by more than 2... 3 ° during amplifier operation;

• spool pressure loss shall be 0.3... 0.6 bar.

When selecting the hydraulic booster pump, they are guided by the following condition:

Required rated pump capacity:

We use a pump with rated capacity:

The diameter of the slide valve is determined by the expression:

3.4 Strength calculation of steering parts

3.4.1 Determination of forces in the steering gear

For strength and durability calculations, it is necessary to determine the forces in the engagement of the piston-rack with the screw in the steering mechanism.

Determine the axial force in engagement, having previously determined the tangent force (see Fig. 3.8):

Axial force on screw Q (see Fig. 39):

3.4.2 Steering shaft strength calculation

We take the material of the steering shaft steel 20.

The steering shaft is calculated by the torsional stress in the hazardous section.

3.4.3 Power calculation of steering gear

The strength calculation of the steering mechanism consists in determining the maximum contact stress for the screw surface of the screw and nut, the permissible axial static load, as well as in calculating the teeth of the sector for bending and contact strength.

First of all, we define the number of balls in the active part of the circulating chain:

From the calculated value of q, at the given radial gap (), according to the graph of dependence of q on at various values ​ ​ of relative radial gap can be determined (the graph is given in lit.1, p. 41, Fig. 20, 14.). It follows from the graph that: Given that the maximum allowable stress, the strength condition is met.

Allowable axial static load:

given in lit.1, p. 41, Fig. 2.16.):

3.4.4 Strength calculation of steering nozzle

We accept 40XN juice material.

The force on the ball pin of the nozzle is:

sections Ι-Ι) and at the same time twists a bipod on a mm shoulder. The equivalent tension stress at the point "" is determined by the formula:

The torsion stress at the point "" is determined by the formula:

where MPa is permissible torsion stress.

Then the torsion stress at the point "":

3.4.5 Strength calculation of longitudinal steering rod

We accept the material of the longitudinal steering rod steel 20.

The longitudinal thrust is calculated for compression and longitudinal bending from the force supplied from the ball pin of the nozzle. Pulling is low hazard.

We define compression voltages of longitudinal steering rod by formula:

3.4.6 Calculation of transverse steering rod

Material of transverse steering rod of steel20.

Figure 3.11 - Diagram for determination of load force of transverse thrust.

The calculation of the transverse steering rod and side levers of the steering trapezoid is made based on the condition that the maximum braking torque is applied to the controlled wheels (Fig. 3.11).

The maximum braking force on controlled wheels is determined by the formula:

Equatorial moment of inertia of cross section of transverse steering rod is determined by formula:

3.4.7 Strength calculation of swivel knuckle lever

We accept material of the side lever steel 40XH.

The pivot lever is calculated by analogy with the steering bow.

Force on ball pin of swivel lever: N.

The force bends the swivel arm on the arm mm (from the ball pin to the dangerous section I-I) and at the same time twists the side arm on the arm mm. The maximum bending stress will be at "" (Fig. 3.9) and the maximum torsion stress will be at "." " The equivalent tension stress at the point "" is determined by the formula:

The torsion stress at the point "" is determined by the formula:

3.4.8 Strength calculation of steering trapezium lever

We accept the material of the steering trapezium lever, 40XN steel.

The steering trapezoid lever is calculated by analogy with the steering bow.

Force on ball pin of steering trapezium lever: N.

The force bends the lever on the arm mm (from the ball pin to the dangerous section I-I) and at the same time twists the lever of the steering trapezium on the arm mm. The maximum bending stress will be at "" (Fig. 3.9) and the maximum torsion stress will be at "." " The equivalent tension stress at the point "" is determined by the formula:

3.5 Selection of main parameters of front axle

The choice of parameters of the front bridge and also parameters necessary for strength calculation is carried out on the basis of a technical characteristic of the MAZ103 bus.

3.6 Calculation of strength of front axle parts

Based on the main dimensions of the front controlled axle, the strength calculation of the front axle beam is made according to the most dangerous of the three main design cases of loading:

- longitudinal force (pushing or braking) reaches a maximum value equal to (no transverse force):

where is the dynamic factor (taken equal to).

3.6.1 Calculation of front axle beam

The front axle beam is calculated for the above three loading cases. Beam sections II, II-II, IIIII are calculated (see Fig. 3.12).

3.6.2 Calculation of beam at bus braking

where ψ - coefficient of redistribution of weight on length;

χ - coefficient of redistribution of weight on height;

Z - relative deceleration (corresponds to maximum value of braking force reaction, Z = 0.8).

The coefficient of mass redistribution by length A is determined by the formula:

The vertical reaction of the soil to the wheels of the Rz1 bus is:

3.6.3 Determination of bending moments

Consider the first calculation case.

Bending moment in vertical plane is determined by formula:

For the second design case, the bending moments for the right and left half of the beam are respectively:

In the third design case, beam sections are checked by bending moment:

The parameters for calculating the elements arising in the front axis beam sections and their values are found in Table 4.

For the third design case, the bending moments are:

The calculation results are summarized in Table 4.

3.6.4 Calculation of geometric parameters of the front axle beam.

The accepted beam has an I-beam section in planes I, II and rectangular in section III. The geometric parameters of the I-beams in the sections are different. Calculation consists in determination of moments of beam resistance to bending in vertical WX and horizontal WY planes, as well as moment of beam resistance to torsion WK in considered sections according to accepted parameters of sections (Table 5). Design diagrams are shown in Fig.3.15.

Results of calculations according to the given formulas for three sections are given in Table 6.

3.6.5 Calculation of stresses in beam sections

Bending stresses in the vertical plane are determined by the formula:

Bending stresses in the horizontal plane are determined by the formula:

Tangent stresses are determined by the formula:

Total bending stresses are determined by formula:

Equivalent stresses from the joint action of bending and torques:

The results of calculations are presented in Table 7.

3.6.6 Calculation of turning fists

Vertical, longitudinal and lateral forces act on the trunnion of the rotary fist. The trunnion should rely on strength, taking into account only bending moments, although it is also affected by compression and tension forces arising from the movement of the bus at a turn, lane change or drift. Stresses from these forces, taking into account the dimensions of the trunnion, are insignificant. Since the wheel hub is mounted on the journal on bearings, the journal itself is not loaded with torque during braking. Torque acts on the rest of the swivel cam.

The most dangerous is the cross section at the point of transition of the trunnion (pin) into the flange or into the body of the swivel cam, located at a distance C along the axis of the wheel from the center of contact of the wheel with the road (Fig. 3.16). In this case, an assumption is made about the vertical location of the front wheels (without camber).

Like the front axle beam, the swivel cam is calculated according to three main loading modes.

The first case of loading. During braking, bending moment, equal to sum of moments in vertical and horizontal planes, acts on trunnion of turning cam:

Then bending stresses for dangerous trunnion section:

Second calculation case. When the bus is driven, bending moments acting on the outer (right) and inner (left) turning knuckles are different. Since the extreme case is considered when the inner wheels have broken off the road, the whole load falls on the outer (right) fist, accordingly, it is advisable to carry out the calculation only for it:

Bending stress, respectively:

Third calculation case. In the vertical plane, a bending moment acts on the trunnion of the rotary cam:

Bending stress:

3.6.7 Calculation of pivots

Pivot lugs and walls of hole for attachment of levers of rotary knuckles are not calculated for strength. Their strength is evaluated during bench and road tests.

Rotary pivot is calculated for bend and cut, and sections of pivot bushing located under support bearing are calculated for crushing. Forces act on the pivot in different planes and directions, causing different stresses. Calculation of the pivot for strength is also carried out for three cases of loading.

The first case of loading. When braking the bus, two vertical forces act on the pivots (relative to the direction of movement of the bus)

Planes - transverse and longitudinal. A bending moment occurs in the transverse plane. This moment is balanced by the moment from the reactions L arising between the pivot and the pivot eyes of the swivel fist:

where a is the distance from the vertical axis of the wheel to the pivot axis,

e is the distance from the horizontal axis of the wheel to the point of action of the force on the upper end of the pivot,

f is the distance from the horizontal axis of the wheel to the point of action of the force on the lower end of the pivot,

As is known, when braking a car at the point of contact of its front wheel with the road, a longitudinal force occurs. When this force is transferred to the center of the wheel axis, we obtain a force acting in the same direction as the bending moment. The force causes reactions F1 and F2 acting between the pivot eyes of the swivel cam and the pivot and directed to the side opposite to the action of the force:

Force is applied at intersection of wheel axis with pivot axis. When the force is transferred in the horizontal plane, a bending moment occurs,

tending to turn the swivel fist around the pivot. A brake support disc is attached to the swivel cam, which during braking also tends to turn the swivel cam. The pivot, holding the swivel knuckle from turning, perceives the following forces:

The direction of action of reaction N2 to the lower pivot eye of the swivel cam coincides with the direction of action of force F2. The reaction N1 acting on the upper eye of the swivel cam is directed to the opposite side to the force F1. Rotation of the swivel cam in the horizontal plane does not occur because the resulting moment is balanced by the moment from the force Q acting on the arm l. Force Q is determined by formula:

As a result of the action of the force Q, reactions Q1 and Q2 appear between the pivot and the pivot eyes of the rotary cam, acting respectively on the upper and lower eyes:

where n and k are the distances from the center of the ball pin attachment of the transverse steering rod to the middle of the lower and upper lugs of the turning cam, respectively;

As a result of taking into account all forces acting on the pivot, we obtain the total reactions Rc and Rn arising respectively between the upper end of the pivot and the upper pivot eye of the turning cam and the lower end of the pivot and the lower pivot eye:

Second case of loading. When driving a car, the directions of action of bending moments from lateral and vertical reactions of the road may either coincide (inner wheel) or not coincide (outer wheel). In the first case, the reaction between the pivot and the pivot eye in the transverse plane:

The calculation is made according to the greatest of forces.

Third calculation case. Under the action of the vertical component of the road reaction to the wheel in the pivot, a bending moment occurs in the vertical plane. This moment is balanced by the bending moment from the reactions arising between the pivot and the pivot eyes of the swivel fist

Process Part

4.1 Introduction

Mechanical engineering is the most important industry. Technological progress in mechanical engineering is characterized not only by the improvement of machine structures, but also by the continuous improvement of their production technology. Its level depends on labor productivity, economical consumption of raw materials and energy resources, quality of output, its durability and reliability of operation. A high level of production technology ultimately leads to cheaper production with the necessary level of quality and reliability.

Course design is an integral part of the course of engineering technology and is a complex work, including the development of the technological process of machining the part, the design of the machine tool, the performance of the necessary technical and economic calculations.

The purpose of this course project is to consolidate, deepen and generalize the knowledge obtained during the study of the course Engineering Technology at lecture, practical and laboratory classes. And also the practically applied use of reference and special literature, standards and standards in combination with theoretical knowledge.

4.2 Part Assignment

The part in question - the steering gear nozzle is part of the steering control of the bus.

One end of part is fitted on splines of steering gear output shaft and fixed on it by means of nut. The other end of the nozzle is connected to the longitudinal steering lever. The main purpose of the part is to transfer the force from the steering mechanism through a longitudinal steering rod to a rotary cam connected to the controlled wheels of the bus. The part is made of 40X steel, the chemical composition and mechanical properties of which are indicated in Table 4.1 and Table 4.2, respectively.

4.3 Part processability analysis

Workability of the part is a set of properties of the product design, which determine its adaptability to achieve optimal costs during production, operation and repair for given qualities, volume of production and conditions of work performance.

The evaluation of the processability of the design can be of two types: qualitative and quantitative. Qualitative assessment characterizes the technicality of the design generically based on the experience of the contractor and is allowed at all stages of design as preliminary. Quantitative evaluation of the workability of the product is expressed by numerical indicators and is justified if they significantly affect the workability of the design in question.

This part is a nozzle made of 40X steel by forging, so the configuration of the outer contour and inner surfaces causes some difficulties in obtaining the workpiece. Forging does not give significant material savings because of which the cost of the product will increase slightly.

Machined surfaces are not difficult to process in terms of accuracy and roughness.

From the point of view of machinability, the part design does not allow the part to be machined in one unit. However, the design and material of the part make it possible to use high-performance processing methods.

Analysis of tolerances for dimensions and surface roughness shows that they are accepted in accordance with operational requirements and do not require the use of special equipment of increased accuracy.

During final processing, it is logical to note that the basing surfaces are preserved throughout the processing of the part, in addition, technological and measuring bases are combined, which reduces the processing error and simplifies further control.

4.4 Choice of production type

The production type according to GOST 3.111983 is characterized by the fixing coefficient of operations Kz.o. = 1 - mass 1 < Kz.o. < 10 - large-scale 10 < Kz.o. < 20 - medium-scale 20 < Kz.o. < 40 - small-scale over 40 - single.

Docking Factor:

Since Kz.o. = 1.5, production is large-scale.

The forms of organization of technological processes in accordance with GOST 14312-74 depend on the established procedure for performing operations, the location of technological equipment, the number of products and the direction of their movement during manufacture, there are two forms of organization of technological processes - group and flow, the main features of which are given in this standard. The decision on the feasibility of organizing an in-line form of production is usually made on the basis of a comparison of the given daily output of products and the estimated daily productivity of the production line with a two-shift mode of operation and its loading on 65... 75%.

4.5 Purpose of machining allowances

Calculation of allowances for machining of the part was selected as per

GOST 750589.

4.6 Calculation and purpose of cutting modes

Consider in detail the purpose of the cutting modes for operation 010 vertical-milling machine mode. 6P13.

Mill surface "G" to maintain the specified dimensions.

Calculation of working stroke length:

4.7 Calculation of technical time standards

The unit-costing time for performing operations in mass production is calculated according to the formula:

Using these formulas, we calculate for the remaining operations, summarize the results into table 4.5.

4.8 Determination of the required quantity of equipment

and plotting its load

For each machine in the process, the load factor and the utilization factor of the machine according to the main time shall be calculated. The machine load factor p is defined as the ratio of the calculated number of machines mp occupied at this process operation to the accepted actual mpr:

Economic part

Introduction

The objective of the economic evaluation of technical solutions is to justify the feasibility of their introduction into the national economy, which consists of analyzing the degree of compliance of machines with their requirements, assessing their advantages over the best of the relevant means of similar purpose and determining the economic effect of introduction into the national economy. The general purpose of the economic evaluation of new equipment is to establish how designed machine structures meet the requirements of high efficiency.

The choice of new machine designs is based on comparison with objects of similar purpose. When establishing the real size of the economic effect as a result of the introduction of new, more progressive machine designs, the models they replace are objects of a similar purpose.

The cost-effectiveness of new machines is estimated on the basis of standard methods and through indicators of cost and capital investments. To obtain an unambiguous solution based on these indicators, it is necessary to measure the compared options of the investment difference and current costs with access to a value assessment of the considered technical solutions.

5.1 Initial Model Data

The basic variant is the control system used on MAZ103.

The projected option - the control system used on MAZ103M.

5.2 Initial Design and Performance Data for Evaluation of Design Competitiveness

The unification indicator will be calculated according to the unification of the parts of the MAZ-103M bus control system:

-unification indicator for the base car:

K = 100% (all parts are unified);

-unification indicator for new car:

5.3 Calculation of costs when manufacturing a designed bus

MAZ-103M

The cost of producing the new MAZ103M bus includes the cost of all units, systems and mechanisms of which the bus consists.

5.4. Calculation of capital investments at the manufacturing stage

new bus (MAZ - 103M).

a car.

5.5. Determining Operating Costs

MAZ bus - 103M

General indicators:

- machine brand (basic): MAZ - 103

- machine brand (analogue in operation): NEOPLAN.

-annual output Ngod = 2000 pcs.

Data required to determine the performance of the new MAZ - 103M bus:

- average technical speed;

- downtime under boarding and disembarking passengers;

- mileage utilization factor;

- average duration of time in order Tc = 14 h;

- load capacity utilization factor;

- vehicle usage factor;

- machine load lifting;

- number of flights per car per year n = 340;

- average length of ride with load L = 200 km;

- machine employment during the year Tgod = 4760 h.

5.5.1 Calculation of performance of the MAZ103M car.

5.5.4 Calculation of operating materials costs (MAZ103M)

Calculated by the formula:

5.5.5 Calculation of maintenance and maintenance costs

(MAZ103M)

Calculation is made according to the formula:

5.5.6 Calculation of tyre costs (MAZ103M)

It should be produced according to the formula:

5.5.7 Calculation of the driver's salary (MAZ103M)

The calculation of the driver's salary with all accruals is carried out according to

From where:

5.5.8 Calculation of overhead costs (MAZ103M)

The calculation of overhead costs is carried out according to the formula:

5.5.9 Calculation of vehicle depreciation costs and overhaul charges (MAZ103M)

It is made according to the formula:

5.5.10 Calculation of capital investments at the consumer of the new car

(MAZ103M)

The total amount of all recorded capital investments per unit of vehicles can be calculated according to the formula:

5.6 Determination of operating costs

NEOPLAN bus

Data required to determine the performance of the analogue bus

5.7Defining the annual economic effect

5.7.1 Calculation of annual economic effect in production

options to be compared:

This value is 810% of the production cost.

5.7 Competitiveness indicators of the designed design

Economic feasibility of investment.

Profitability for profit remaining for the needs of the enterprise according to the designed version is determined by the formula:

Conclusion

As a result of the completed economic part of the diploma project, the costs at the stage of production and operation of the MAZ103/NEOPLAN bus were determined. In addition, the work gives an economic assessment of the technical decisions made during the project. The price of the designed bus, conditions of competitiveness and performance indicators were also determined.

Literature

1. Gainutdinov E.M. Methodological guidelines for the implementation of the economic section of the diploma project and coursework for students of engineering specialties - Mn.: BGPA, 1995. 24 pages.

2. Calculation of cost-effectiveness of new equipment: Handbook. - 2nd ed., Redesign. and additional/Ed. K.M. Velikanov. -L.: Mechanical Engineering, Leningr. Division, 1990.

3. Economics of the enterprise: Textbook for economic universities, -

Iz1.1y, reslave. and additional/Under the general editorial A.M.Rudenko. -Mn. 1995,-

475 pages.

4. Legislation of the Republic "Belarus on taxation (as of the time of calculations in 2005).

5. BabukI.M. Methodological guidelines for calculating the cost-effectiveness of introducing new technological processes for students of engineering specialties (diploma design), - Mn.: BGPA, 1992. 18 pages.

1 Course design of machine parts: Right. Allowance. Part 2/

[A.V. Kuzmin, N.N. Makeichik, V.F. Kalachaev, etc.]. - Mn.: Vysh. school, 1982. -334 p.: il.

2 Cars: Design, design and calculation. Control systems and running gear: Textbook for universities/A.I. Grishkevich, D.M. Lomako,

V.P. Avtushko and others; Ed. A.I. Grishkevich. - Mn.: Vysh. shk., 1987. -200 s.: il.

3 Raympel J. Chassis: Steering/Per. with German V.N. Palyanova; Ed. A.A. Galbreich. -M.: Engineering, 1987. -232 p.: il.

4 Conception of lectures on the discipline "Design of car control systems."