RU2151936C1 - Hydrodynamic variable-speed gearbox - Google Patents

Abstract

FIELD: automobile manufacture; mechanical engineering. SUBSTANCE: gearbox includes primary and secondary shafts located coaxially and connected with engine flywheel and cardan shaft, torque hydrodynamic converter located in case rigidly connected with engine cylinder block; arranged in closed circulation volume of hydrodynamic converter are primary cylinder of pump and additional cylinder of pump rigidly fitted on primary shaft, reactor cylinder of additional torque converter located around additional primary cylinder of pump, reactor cylinder of main torque converter located around primary cylinder of pump and around reactor cylinder of additional torque converter, secondary cylinder turbine rigidly connected with secondary shaft forming closed cylindrical space around main torque converter, additional torque converter, primary cylinder of pump and additional primary cylinder of pump; main and additional torque converters are provided with respective control rods. EFFECT: facilitated manufacture; enhanced durability and efficiency. 11 cl, 37 dwg, 1 tbl

Description

 The invention relates to the field of automotive, mechanical engineering and can be used in transmissions of both front-wheel drive and rear-wheel drive cars, trucks, and other vehicles.

 Known hydrodynamic automatic transmission (AKP), German patent: DE 3263522, published August 2, 1966 N.T. GENERAL DUAL RANGE TORQUE CONVERTER TRANSMISSION and German patent DE 2910111 C2, cl. F 16 H 41/04 of 10/17/1991, which consists of a hydraulic clutch, the principle of which is based on “slipping”, and an automatic transmission, consisting of several simple or complex planetary gears installed in rows (planetary gears), the control of which occurs automatically with the help of hydrodynamic devices (oil pump and valve system) that control friction brake bands that stop, or, conversely, allow the control discs to rotate. The control disk can be dressed on any of the gears: crown gear, carrier, sun gear, depending on the inclusion of a simple planetary system.

Also known is a transmission comprising a housing and a CVT and gearbox located therein (German patent N 4104487, class 60 K 17/06, 1991). Also known are:
1. The continuously variable transmission of the vehicle, comprising a housing, a sequentially located in it and connected by a variator and a reverse gear (RF patent RU 2108926 C1, CL 60 K 17/08, 09/05/1998).

 2. Bus hydromechanical transmission, comprising an engine, input and output shafts, a hydrodynamic torque converter combined with a hydraulic retarder, in the circulation circle of which are pump, turbine and reactor wheels, a brake retarder and a planetary gearbox including planetary gears and friction controls (RF patent RU 2104431 C1, class F 16 H 47/06, 05/20/1998).

 3. A composite transmission gearbox comprising a multi-speed main transmission section connected in series with a multi-speed additional transmission section, an input shaft extending into the main transmission section, a main shaft extending from the main transmission section to the additional transmission section, and an output shaft extending from the additional section transmissions (RF patent RU 2104171 C1, class B 60 K 17/08, 05/20/1998).

 The disadvantages of these transmissions is that they are very bulky, have a lot of weight, require a gearbox that includes planetary gears and friction controls, and V-belt variators have a limited range of gear ratios, and the disadvantage is the complexity, laboriousness and high cost of the product . The presence of friction elements (brake bands) in the hydrodynamic control of the box determines the low operational capabilities of the product as a whole.

The proposed device is a continuously variable hydrodynamic gearbox installed between the engine and the vehicle’s transmission instead of the clutch and gearbox (mechanical or hydrodynamic automatic gearbox) and is their full functional substitute, but with additional operational capabilities and consumer properties, such as:
I ^* smooth adjustment of the gear ratio of rotation from the engine to the transmission in any possible numerical values of the gear ratio within specified limits with the following capabilities:
- electrical control of the entire range of gear ratio,
- control using an electronic control unit for the entire range of transmission coefficient,
- control using a radio-controlled signal over the entire range of transmission coefficient,
- automatic control of the entire range of gear ratio,
- and all possible combinations thereof;
II. ^* anti-theft system, not as a separate hinged structure, but as an integral part of the car, inscribed in the design of a hydrodynamic continuously variable transmission.

III. ^* forced automatic control using a radio-controlled signal speed of the car (cars) for:
- stopping and subsequent detection of stolen vehicles, if the vehicle at the time of its abandonment by the owner was unencoded for any reason,
- regulation of the maximum permissible speed of the car (automobiles) on the roads in places with increased danger by radio means of the traffic police;
IV. simplicity of design, low cost and durability, since the hydraulic drive has completely no friction transmission of rotation from the primary shaft to the secondary and almost completely no gear transmission of rotation from the primary shaft to the secondary, 90% of metal-intensive parts have the opportunity to use selumin casting, and a complete rejection the use of friction parts and friction brake bands in the design makes the design more durable compared to the hydrodynamic automatic gearbox before h (ACT);
V. higher efficiency of a hydrodynamic stepless gearbox as compared to the efficiency of a hydrodynamic automatic gearbox, the hydro clutch of which uses the “slippage” effect to achieve “softness of travel”, or a variation of the “slippage” effect with automatically engaged friction clutches to increase the efficiency. The efficiency of a hydrodynamic stepless gearbox is higher than the efficiency of a hydrodynamic automatic gearbox due to the refusal to use the design of the hydrodynamic stepless gearbox of the “slippage” effect at high engine speeds and minimizing it at low engine speeds.

 Preliminary calculations show that a continuously variable hydrodynamic gearbox can fit into the overall dimensions of existing structures and, therefore, does not require a revision of the vehicle design. It can also be installed on a pre-owned car by replacing the clutch and gearbox with a hydrodynamic continuously variable gearbox and the subsequent re-equipment of the passenger compartment.

The objectives of the invention are:
I. Having preserved the advantages of the existing prototype, create a simpler, more reliable, technologically advanced hydrodynamic device (hydraulic drive), the transmission coefficient of which changes steplessly in the forward and reverse directions of movement:
- symmetrically
- asymmetrically,
overall dimensions and weight of which do not depend on the required gear ratio. Stepless hydraulic drive can be used in vehicles:
1) with an internal combustion engine,
2) with a high-speed low-power direct current electric motor with a very massive flywheel mounted on an electric car: a low-power high-speed direct-current electric motor, when the batteries are turned on, spins up a very massive flywheel up to 15,000 - 20,000 rpm and maintains this angular speed constantly, but for the dynamics of the electric car corresponds to a stepless hydraulic actuator, small overall dimensions, weight and high gear ratio (from 1000 to 1) which allow:
a) effectively and dynamically control an electric vehicle, despite the low power of the electric motor,
b) to optimize the energy consumption of battery batteries and increase the range of the electric vehicle between battery charges,
3) with a fuelless engine, if it is possible to create a more powerful and high-speed design, while facilitating the control of turning on and off the fuelless engine.

 The reverse gear is turned on in all cases by turning on the reverse gear button, and the gear ratio of this hydraulic drive is controlled by electric control.

II. Based on the electrical control of the transmission coefficient, create:
1) control using an electronic control unit for the entire range of transmission coefficient,
2) control using a radio-controlled signal over the entire range of transmission coefficient,
3) automatic control of the entire range of gear ratio.

 III. Based on the control using the electronic gear ratio control unit, create an anti-theft system as an integral part of a hydrodynamic stepless gearbox installed inside the body of a hydrodynamic stepless gearbox.

 IV. Based on the control using a radio-controlled signal over the entire range of the transmission coefficient, create a forced automatic control using a radio-controlled signal from the outside speed of the car (reducing the vehicle speed to the level of the maximum allowable value in places with increased danger).

 These tasks are achieved by the fact that the hydrodynamic stepless gearbox, which controls the beginning and end of the vehicle’s movement, “forward-backward” movement, steplessly controls the gear ratio of rotation from the engine to the driveshaft, contains primary and secondary shafts located coaxially and connected, respectively, with engine flywheel and propeller shaft, hydrodynamic torque converter located in a housing rigidly connected to the engine block, in a closed volume During the circulation of the hydrodynamic torque converter, the primary pump cylinder and the additional pump cylinder are rigidly mounted on the input shaft, the reactor cylinder of the additional torque converter located one around the other primary pump cylinder, and the reactor cylinder of the main torque converter located around the pump primary cylinder and around the reactor cylinder additional torque converter, turbine of the secondary cylinder, rigidly connected to the secondary shaft forming a closed cylindrical space around the main torque converter, the additional torque converter, the primary pump cylinder and the additional primary pump cylinder, while the main and additional torque converters are made with the corresponding control rods.

 The main feature that determines the solution of the tasks set is the principle of forced rotation of the torque converter and the principle of forced movement of the torque converter along its axis of rotation, in contrast to a rigidly mounted torque converter between a rotating drive disk of a hydraulic pump and a driven disk of a hydraulic turbine in hydraulic clutch devices. For the implementation of the principle of forced displacement and forced rotation of the torque converter, the cylindrical shape of the leading hydraulic pump and the driven hydraulic turbine with blades located on the working cylindrical surfaces is the most acceptable form, the end (disk) sides of the cylinders in this case turn out to be inoperative. In this regard, the shape of the torque converter must be cylindrical.

 The control of the rotation of the torque converters and the movement along the axis of rotation is carried out using stepper motors. This allows you to implement electrical control of the hydraulic drive, control using an electronic control unit, automatic and hydraulic control using a radio-controlled signal.

 The closed internal (working) cavity of the driven hydraulic turbine allows using the oil pump to collect the working fluid displaced through the bearing and rod clearances and to control the pressure in the working cavity of the driven hydraulic turbine. By increasing the pressure in the working area, you can increase the efficiency of the hydraulic drive.

In the closed circulation volume of the hydrodynamic torque converter are located:
- sections of the primary cylinder of the pump and sections of the additional primary cylinder of the pump, rigidly mounted on the input shaft,
- sections of the reactor cylinder of the additional torque converter, sequentially alternating left and right, located around the sections of the additional primary pump cylinder, having two support points on the input shaft through sliding bearings, and also having the ability to move "left-right" along the axis by a distance equal to the width of the section , using rods passing along a sliding fit through a torque converter sleeve and rigidly fixed to the left section of the reactor cylinder of an additional torque converter, tulkus converter rigidly secured to the housing of the hydrodynamic continuously variable transmission, so stocks, cylinder reactor with an additional converter, have no possibility of rotation around the symmetry axis of the system in a dynamic mode,
- sections of the reactor cylinder of the main torque converter, sequentially alternating left and right, located around sections of the primary cylinder of the pump, connected at the ends with a turbine forming a closed cylindrical space around sections of the reactor cylinder of the additional torque converter, having two support points, on the input shaft and on the torque converter sleeve plain bearings, as well as having the ability to move "left-right" along the axis of rotation to the width of the section using the second pair s rods passing along the sliding fit through the torque converter bushing and connected to the left side of the torque converter’s turbine cylinder of the main butt-end torque converter through a rolling bearing, and the rotation speed and direction of rotation of the main torque converter around the axis of rotation of the system in dynamic mode are determined by the relative position of the sections of the additional converter the primary cylinder of the pump and sections of the reactor cylinder of the additional torque converter.

 Changing the direction of rotation of the main torque converter, from free rotation to one side to forced rotation to the other side, determines the beginning and end of rotation of the secondary cylinder, which forms a closed cylindrical space around the main torque converter, additional torque converter, primary pump cylinder, and additional primary pump cylinder. The turbine of the secondary cylinder is rigidly connected to the secondary shaft. The direction of rotation of the turbine of the secondary cylinder in dynamic mode is determined by one of the two extreme positions of the rotating main torque converter, which are fixed by a second pair of rods and a return spring.

 The transmission coefficient of rotation from the primary cylinder of the pump to the secondary cylinder of the turbine is determined by the ratio of the radii of the secondary and primary cylinders and the speed of the forced rotation of the main torque converter. The beginning and end of the vehicle’s movement, the forward-backward movement, and the transmission coefficient of rotation from the engine to the driveshaft are controlled by linear movement along the axis of rotation of the primary and secondary shafts of two pairs of rods. A pair of rods connected to the main torque converter, one of its two extreme positions, determines the movement of the vehicle either forward or backward, and a pair of rods connected to an additional torque converter that controls the free and forced rotation of the main torque converter, as well as the speed of the forced rotation of the main torque converter, determines the beginning and end of the vehicle’s movement, and also linearly changes the transmission coefficient of rotation from the engine to the driveshaft.

Since the linear displacement of the rods and the forces that prevent the linear displacement of the rods are small, the rods can be controlled using low-power motors or stepper motors (stepper motors are preferable) located outside the hydraulic actuator housing or inside it (placement of stepper motors outside the housing is most preferable ) Electrical gear ratio control,
the beginning and end of the movement, the movement back and forth, makes it possible to provide:
- management using an electronic control unit,
- automatic control,
- control using a radio-controlled signal and all possible combinations thereof, as well as, on the basis of electronic control of the beginning and end of the movement, the gear ratio of rotation, back and forth movement, the anti-theft system, as an integral part of the vehicle’s structure, inscribed in the design of a continuously variable hydrodynamic gearbox, unlike the well-known individual hinged structures. On the basis of control using a radio-controlled signal, forced automatic control using a radio-controlled signal of the vehicle speed to stop and subsequently detect a stolen vehicle, if at the time it was left by the owner for some reason turned out to be unencoded.

 Based on control using a radio-controlled signal, regulation of the vehicle speed to the level of the maximum permissible speed, if the car exceeds it on roads in places with increased danger, in places with limited speed, by radio.

 An additional oil pump with a gear drive of rotation from the input shaft, located inside the hydraulic drive housing between the housing and the turbine of the secondary cylinder, allows not only to collect and “return” to the working area the oil squeezed out by the working pressure in dynamic mode through the sliding bearings of the torque converter sleeve and the gaps between the rods and bores of the torque converter sleeve, but also significantly increase the pressure in the working area. The pressure level in the working area is controlled by a pressure reducing valve. The pressure in the system is limited by the rigidity of the turbine design of the secondary cylinder. Initial increased pressure in the working area makes it possible to increase the efficiency of a hydrodynamic continuously variable gearbox in comparison with the analogue.

The design of a continuously variable hydrodynamic gearbox is illustrated by drawings, where in FIG. 1 shows the design of the torque converter; in FIG. 2 - a flat configuration of the torque converter blades; in FIG. 3 - wedge-shaped configuration of the torque converter blades; in FIG. 4a - a hydraulic cylinder; 4b - hydraulic cylinder with torque converter blades located inside; FIG. 5 - torque converter is inhibited; FIG. 6 - torque converter is not inhibited; FIG. 7 - the design of the hydraulic drive without the possibility of changing the gear ratio; FIG. 8 - design of a hydraulic actuator with the ability to change the gear ratio, FIG. 9 shows growth characteristics of the transmission coefficient for flat and wedge-shaped blades, FIG. 10 - hydrodynamic continuously variable gearbox; FIG. 11 - additional torque converter; FIG. 12 - a primary cylinder and an additional primary cylinder; FIG. 13 - the right section of the main (additional) torque converter;
FIG. 14 - left section of the main (additional) torque converter; FIG. 15 - left (right) plate of the main (additional) torque converter; FIG. 16 - torque converter of the secondary cylinder; FIG. 17 - the primary cylinder of the pump; FIG. 18 - an additional primary cylinder of the pump; FIG. 19 - additional torque converter; FIG. 20 - a cover of the reactor cylinder of an additional torque converter; FIG. 21 - the right section of the reactor cylinder of the additional torque converter; FIG. 22 - left section of the reactor cylinder of the pump additional torque Converter.

 The continuously variable hydrodynamic gearbox contains a torque converter sleeve 1 (see Fig. 10), a main torque converter 2 (Figs. 7, 8, 10), a primary shaft 3, a primary 4 and a secondary 5 cylinders, bearing rings 6, a rod 7, a secondary shaft 8 , an additional primary cylinder 9 and an additional torque converter 10, an engine 11, a flywheel 12, an oil pump 13, stepper motors 14, a housing 15, a driveshaft 16, gears of an oil pump drive 17.

When the ignition is turned on, the engine 11 (see Fig. 10) starts to rotate the flywheel 12. The flywheel rotates the input shaft 3, on which the primary cylinder of the pump 5, the additional primary cylinder of the pump 9 and the drive gears also begin to rotate oil pump 17. The rotation of the gear wheel drive the oil pump ensures the presence of working fluid pressure in the working area of the solar hydraulic actuator, and the oil pump valve provides the desired value of the working pressure P _slave .

 The stepper motor 14, connected through the bearing by the rod 7 to the main torque converter 2, has two forward-backward operating positions and is switched on by a forward-backward driver button depending on the desired direction of movement.

 A stepper motor 14, rigidly connected by a rod 7 with an additional torque converter 10, kept from turning rigidly fixed by brackets to the housing 15 by a torque converter sleeve 1, is responsible for the linear movement of the additional torque converter along the axis of rotation of the system, thereby creating all possible options for overlapping sections of the additional primary pump cylinder 9 with left (Fig. 22) and right (Fig. 21) sections of the additional torque converter 10, thus providing speed control I and the direction of rotation of the main torque converter 2.

The rotating working sections of the additional primary cylinder of the pump 9, combined with the left or right sections of the additional torque converter 10, disturb the working fluid under the working pressure P _slave. .

 The divergence of the flow perturbed by the blades of the sections of the additional primary cylinder of the pump 9 fluid, directed by the blades of the sections of the additional torque converter 10 to the desired left (for moving backward) or right (for moving forward) side, begins, depending on the location of the sections of the additional torque converter and the additional primary pump cylinder, or slow down the rotation of the main torque converter 2 (in the neutral position, it rotates in the direction opposite to the rotation of the primary shaft), which corresponds to the beginning of rotation of the secondary shaft 8 in the direction of rotation of the primary shaft in a lower gear, or to stop the rotation of the main torque converter (higher lower gear), or to rotate the main torque converter in the direction of rotation of the primary shaft, which corresponds to the rotation of the secondary shaft in higher gears.

The rotating working sections of the primary cylinder of the pump 5, combined with the left (for moving backward) or with the right (for moving forward) sections of the main torque converter 2, perturb the working fluid under the working pressure P _slave. . The divergence of the flow of the fluid disturbed by the blades of the sections of the primary pump 5 and directed by the blades of the sections of the main torque converter 2 to the desired left (for moving backward) or right (for moving forward) side starts to rotate in the right side of the turbine blade of the secondary cylinder 4, to which the output ( secondary) shaft 8. The rotation speed of the secondary shaft 8 depends on the rotation speed of the primary shaft 3 and on the overlap area of the sections of the additional torque converter 10 with the working sections of the additional primary 9 of the pump cylinder.

 1. Work on the creation of automatic transmissions began in 1932.

Different designs were offered:
the movement of gears and synchronizers in the appropriate position using fluid pressure, compressed air pressure, electricity, air discharge. Chrysler, Chevrolet, Packard firms used vacuum to move gears and synchronizers. The company Cutler-Hammer on Premier cars made the first devices on electric solenoids: there were 4 buttons on the instrument panel, the driver selects 1, 2, 3, R-rear. desired gear, engages the clutch, then engages the gear with a button. The lack of solenoids is their high power consumption.

 In contrast to this system, in the “electric arm” used later by Hudson, a small solenoid moved the valve, which shifted gears under the action of vacuum. The American-made Dodge was the first to use a fluid coupling in combination with a conventional manual gearbox. Advantages: the engine works without overloads when towing heavy loads due to slip of the coupling, the coupling absorbs vibration and unevenness of rotation of the crankshaft, protects the engine from shock loads, also due to slip of the coupling you can keep the car "on the hill" with the engine on with the gas pedal, not including brakes.

 Modern automatic transmissions began to be produced after successful use in the construction of a simple planetary system. Possible options for the operation of simple planetary gears 8.

 In planetary systems, in contrast to the gears of conventional gearboxes, where almost any gear ratio can be implemented, depending on the number of gear teeth engaged, the possible number of gear ratios is limited. In one planetary gearbox it is impossible to get even those options that are reflected in the table, since it is impossible to change the places of connection. Therefore, in order to get a three or four-stage planetary gearbox, it is necessary to install several planetary gears or apply systems with two satellite sets. In modern boxes, both of these methods are applied.

II. The principle of hydraulic clutch operation is based on the transmission of rotation from the primary disk by means of a medium (oil) perturbed by the blades of the primary disk to the secondary disk, also equipped with blades. To increase the hydraulic coupling efficiency, between the rotating primary and secondary disks (ω ₁ ≠ 0; ω ₂ ≠ 0; ω ₁ > ω ₂ ) a torque converter is rigidly fixed to the hydraulic clutch housing (ω _tr = 0 is always for hydroclutch),
where: ω ₁ is the angular velocity of the primary disk;
ω ₂ is the angular velocity of the secondary disk;
ω _Tr - the angular velocity of the torque Converter (see Fig. 1).

The torque converter, passing through its blades the medium (oil) perturbed by the blades of the primary disk, orientates its movement in space in such a way that it creates a “normal” (at a perpendicular angle) liquid drop on the secondary disk blades, thereby increasing the efficiency of the entire assembly. In reality, for hydrolinking, ω ₂ <ω ₁ ≈ 20% due to elastic losses in the fluid (slippage effect). The slippage effect decreases in the case of an increase in ω ₁ and increases in the case of a decrease in ω ₁ , thereby actually ensuring the “softness” of the stroke of the entire system, and determines the adhesion of the primary and secondary shafts starting from a certain value of ω ₁ (usually not lower than 1000 rpm ) Slippage decreases to a certain level with an increase in ω ₁ and in the region of 2000–3000 rpm, the decrease in slippage stops and then it ≠ 0.

Theoretically, slippage can be reduced to zero due to the phenomenon of “hard water” (term), that is, to increase the oil pressure in the system, but then:
1. With an increase in efficiency, the smoothness of the stroke decreases (for the sake of which, in fact, hydro clutch is made).

 2. It will require the installation of a second shell of the hydraulic clutch housing and the installation of an oil pump between two housing shells. That is, the futility of this path is obvious.

 All modern automatic transmissions, having appreciated the advantages of the Dodge, are equipped with hydraulic clutch to improve the dynamic characteristics of the car (smoothness). Although the hydro clutch reduces the efficiency of the car, but anyone who can afford such an expensive device (automatic gearbox + hydro clutch) can also compensate for the lost hydro clutch power due to a more powerful engine.

 III. The principle of operation of the planetary mechanism is that the rotation from the gear of the primary shaft (inner gear) is transmitted to the gear of the secondary shaft (outer gear) by the gears of the carrier (there are at least three of them) only if the gears of the carrier are inhibited. The gear ratio in this case is equal to the ratio of the circumference of the two gears. The direction of rotation of gears 1 and 2 is opposite (see table 1 case 3).

 Let’s try to imagine a model of a planetary mechanism in which both gears are made not in a flat form, as usual, but in the form of cylinders, and instead of teeth they have blades: the outer cylinder has the inner cylinder and the inner cylinder has the outer one, the role of the gears (motion transmitter from the primary gear to the secondary) performs an oil medium (see Fig. 1, section A-A through a torque converter).

 Unlike the planetary mechanism, the direction of rotation of the inner master cylinder and the outer follower is the same, but unlike the hydraulic clutch, the gear ratio of which is w 1, y of the hydraulic drive, the gear coefficient is theoretically for the case of “hard water” and inhibited torque converter equal to the ratio of the circumference or the ratio of the outer radii and the inner cylinders, or the square of the ratio of their radii (we consider these two cases below; with different designs of the torque converter, the patterns will be different).

 Case 1

The torque converter blades have a flat configuration. The rotation of the fluid flow occurs with a change in volume (that is, h _blades - along the entire length of the const). see FIG. 2.

The second law of hydrodynamics (the law of conservation of momentum for elastic media)
I ₁ • ω ₁ = I ₂ • ω ₂ + I _x • ω _x ,
where I _x • ω _x are elastic losses.

We write this law for the case when I _x • ω _x = 0, that is, for the case of the absence of elastic losses ("hard water").

I ₁ • ω ₁ = I ₂ • ω ₂
The transmission coefficient k = ω ₁ / ω ₂ , therefore
k • I ₁ = I ₂ , i.e. k = I ₂ / I ₁ .

Сила F = P • S, где P - давление жидкости, S - площадь цилиндра, (S₁ - площадь внутреннего цилиндра, S₂ - площадь наружного цилиндра).The moment of inertia is I = F • r, therefore

Force F = P • S, where P is the fluid pressure, S is the area of the cylinder, (S ₁ is the area of the inner cylinder, S ₂ is the area of the outer cylinder).

для замкнутого контура давление во всех точках одинаково, следовательно,

если l₁ = l₂, то

. Для случая r₂ = 2 • r₁, k = 4.From here

for a closed circuit, the pressure at all points is the same, therefore,

if l ₁ = l ₂ then

. For the case r ₂ = 2 • r ₁ , k = 4.

при том, что внутренний цилиндр вращается с угловой скоростью ω₁ = Δr₁/r₁•t. Коэффициент передачи вращения k = ω₁/ω₂ = 2•Δr₁/Δr_x, см. фиг. 4б.In confirmation of this, we consider another case. How much does the radius of the outer cylinder change if the radius of the inner cylinder has changed (r ₁ + Δr), and the volume of fluid between the two cylinders remains - const (see Fig. 4a). The pressurized cylinder A is located inside the pressurized cylinder B, and the intermediate volume between the two cylinders is completely filled with ideal fluid. Consider the case when the radius of cylinder B is equal to the two radii of cylinder A, i.e. r ₂ = 2 • r ₁ . If somehow the volume of cylinder A has changed due to a change in radius (r ₁ + Δr), then how much will the radius of cylinder B change? That is, what is Δr _x equal to?
If subsequently, in the volume occupied by the liquid, place the torque converter blades L ₃ (Fig. 4b) to change the direction of the detonating liquid: then the outer cylinder will rotate towards the turn of the blades L ₃ by an angle φ, approximately equal to the ratio Δr _x to the radius of the cylinder B, those. to 2 • r ₁ . If, in practice, a change in the volume of cylinder A is carried out by rotating it around the axis of symmetry with a large angular velocity, and equip its outer surface with blades L ₁ , then the blades will, together with cylinder A, “squeeze out” the liquid from the region r ₁ + Δr. If the "angle of attack" at the blades L ₁ is equal to 45 ^o , then the vector velocity of the elementary volume of the fluid that "bounced" from the blades L ₁ will be equal in numerical expression to the vector speed of the blades L ₁ , the directions of the vectors of their velocities will be perpendicular. The "elementary volumes" of the detonated liquid, collected in a common stream and directed in the right direction by the blades of the torque converter L ₃ , interact with the inner side of the outer cylinder, also equipped with receiving blades of L ₂ . The outer cylinder, as mentioned above, begins to rotate at an angular speed

despite the fact that the inner cylinder rotates with an angular velocity ω ₁ = Δr ₁ / r ₁ • t. Rotation transfer coefficient k = ω ₁ / ω ₂ = 2 • Δr ₁ / Δr _x , see Fig. 4b.

отсюда

тогда π•(4•r₁•Δr_x+Δr $\genfrac{}{}{0ex}{}{\text{2}}{\text{x}}$ -2•r₁•Δr₁-Δr $\genfrac{}{}{0ex}{}{\text{2}}{\text{1}}$ ) = 0.
внесем +4 • r₁ ² - 4 • r₁ ²,
получим 4•r $\genfrac{}{}{0ex}{}{\text{2}}{\text{1}}$ +4•r₁•Δr_x+Δr $\genfrac{}{}{0ex}{}{\text{2}}{\text{x}}$ -4•r²-2•r₁•Δr₁-Δr $\genfrac{}{}{0ex}{}{\text{2}}{\text{1}}$ = 0
то есть

тогда

но,

Теперь, если, например, r₁ = 10, а Δr₁ = 1 (напомним, что это случай r₂ = 2 • r₁), то k ≈ 4.S _c = π • (2 • r ₁ ) ² -π • r $\genfrac{}{}{0ex}{}{\text{2}}{\text{1}}$ = 3 • π • r $\genfrac{}{}{0ex}{}{\text{2}}{\text{1}}$ .
hence

from here

then π • (4 • r ₁ • Δr _x + Δr $\genfrac{}{}{0ex}{}{\text{2}}{\text{x}}$ -2 • r ₁ • Δr ₁ -Δr $\genfrac{}{}{0ex}{}{\text{2}}{\text{1}}$ ) = 0.
we introduce +4 • r ₁ ² - 4 • r ₁ ² ,
we get 4 • r $\genfrac{}{}{0ex}{}{\text{2}}{\text{1}}$ + 4 • r ₁ • Δr _x + Δr $\genfrac{}{}{0ex}{}{\text{2}}{\text{x}}$ -4 • r ² -2 • r ₁ • Δr ₁ -Δr $\genfrac{}{}{0ex}{}{\text{2}}{\text{1}}$ = 0
i.e

then

but,

Now, if, for example, r ₁ = 10, and Δr ₁ = 1 (recall that this is the case r ₂ = 2 • r ₁ ), then k ≈ 4.

Case 2. ^*
The torque converter blades are wedge-shaped. The rotation of the fluid flow is without changing its volume. Size d is the distance between the blades along the entire length - const (see Fig. 3).

In this case, the fluid in the torque converter simultaneously interacts (when the inner cylinder rotates, of course) with the same areas (see Fig. 5), and not with different areas (as in case 1), therefore, the transfer coefficient in this case is calculated as in simple planetary system (table 1, case 3) k = r ₂ / r ₁ . For the case r ₂ = 2 • r ₁ , k = 2. Conclusion: for different forms of torque converter blades, the transmission coefficient is considered differently.

 So, we have seen that for a hydraulic actuator for the case of a closed oil volume and a braked torque converter, the transmission coefficient is a constant value that is not equal to 1 as in hydraulic coupling and can be systematically made any way, say, equal to 4 or 2, etc.

Now we consider the case when ω _tr ≠ 0.
Case 3

When the torque converter is not inhibited, transmission of rotation from the primary (internal) cylinder to the external one does not occur (this is clear, since there is an analogy with the solar mechanism in the case of an unbraked carrier - the carrier begins to rotate). The torque converter begins to rotate in the direction opposite to the turning side of the torque converter blades with an angular velocity ω of free rotation, under the action of forces arising at the moment of changing the direction of the liquid when it moves along the torque converter blades (see Fig. 6). The average point of application of these forces is in the region of the “knee” of the torque converter blade, respectively, the moment of the forces accelerating the torque converter is considered from these conditions:
M _arrears = ∑ F _{1 ... n} • r _{mid knee} ,
where r is _{average knee} - the distance from the center to the average point of application of forces ∑ F _{1 ... n} . Quantitatively, this moment of forces is small, since on the radius of the bend of the knee, the addition of forces occurs vectorly, and if you need to further reduce this moment, then you can distance r _{average. of the knee} (the distance from the center of rotation of the torque converter to the center of application ∑ F _{1 ... n} forces) to reduce, approaching the center of rotation the start of bending of the knee. Optimal r _{avg. knee} ≈ 2/3 r.

That is, if it is necessary to begin the movement of the secondary cylinder, it is enough for us to apply a braking force to the rotating shaft of the torque converter, comparable in magnitude to M _decomp. but directed in the opposite direction. If you apply the braking force “softly” and adjust its value from 0 to M _max. Starts a "soft" acceleration of the secondary cylinder, ω ₂ varies from 0 to a value ω _1/4 to FIG. 6, and ω ₂ varies from 0 to a value of ω _1/2 for FIG. 5 for different torque converter designs with different speeds. For simplicity, we will always consider the case r ₂ = 2 • r ₁ . Soft acceleration is provided by soft braking and slippage. If slippage due to “hard water” is excluded (raising the oil pressure), then the softness of the acceleration of the secondary cylinder will be determined only by the softness of braking of the torque converter, and slipping losses (as in the hydraulic clutch system) can already be eliminated in principle, therefore, raising the pressure in the hydraulic drive system , we get the transmission of rotation from the primary cylinder to the secondary one with high efficiency (without loss of slippage), with the possibility of "soft" control of the beginning and end of rotation of the secondary cylinder and with lowering it with a transmission coefficient of rotation (4 or 2) depending on the design of the torque converter.

 Case 4

The case of forced rotation of the torque converter in the direction opposite to the direction of free rotation of the unbraked torque converter. To consider this case, it is most convenient to consider the design of a torque converter with wedge-shaped blades (case 2). The interaction of the primary cylinder with the secondary occurs on the circumference of the secondary cylinder (the interaction, of course, does not occur on the circumference, but on the areas equal to the product of the circumference by the length of the cylinders. We simplify this because these cylinder-in-cylinder designs have their lengths approximately equal), equal to l ₂ = (nl) / d, where n is the number of torque converter blades, ad is the distance between them (see Fig. 3). In this case, structurally l ₁ ≈ l ₂ , therefore, while the inner cylinder makes one revolution, the outer (r ₂ = 2 • r ₁ ) will make half a revolution. If under these conditions we will force the torque converter to rotate forcibly with a speed of ω _1/2 , then during this time the interaction of the inner and outer cylinders will occur along the entire circumference of the inner cylinder and the outer cylinder, and therefore, it will simultaneously make both the inner and outer cylinder. That is, in this case, the transmission coefficient will be equal to 1.

The picture will be as follows: the inner and outer cylinders rotate with the same angular velocity, provided that the torque converter cylinder rotates forcibly in the same direction with the angular velocity ω _tr = ω _1/2 . Naturally, if the forced angular rotational speed of the torque converter ω _tr. will be less than ω _1/2 , but greater than 0, the transmission coefficient will be greater than 1, but less than 2. That is, by inhibiting the free rotation of the torque converter (see case 3) and subsequently forcibly rotating it (see case 4), ω _tr. = _{-ω tr.} ÷ + ω _1/2 , we can smoothly change the transmission coefficient from ∞ to 1. At the time when ω _tr. = 0, the transmission coefficient for r ₂ = 2 • r ₁ for wedge-shaped torque converter blades is 2. For r ₂ = 2 • r ₁ and the design of the torque converter with flat blades, the transmission coefficient also varies from ∞ to 1, the only difference is that when ω _tr. = 0, the transmission coefficient for flat blades is 4, that is, the steepness of the growth characteristics of the transmission coefficient for flat blades is higher than for wedges (see Fig. 9). All this could not be explained for a long time, but just remember the game of tennis: you hit the ball with a racket flying towards you, trying to change the direction of its flight by 90 ^o . The direction of the ball’s flight changes, and its flight speed becomes equal to the vector sum of the speeds of the ball and the racket during a strike.

 The design of the hydraulic drive without the possibility of changing the gear ratio (see Fig. 7). A closed oil volume creates a secondary cylinder 4, the right part of which is connected to the shank of the universal joint, and the left one slides along the torque converter shaft through a sliding bearing 6. The inner diameter of the torque converter sleeve 1 sits on the primary shaft 3 through a sliding bearing. For the case of "wet water" (lack of initial pressure in the system), both shafts are equipped with grease seals (similar to hydraulic clutch gaskets) and oil from the system will not leak.

 The design of the hydraulic drive with the ability to control the transmission coefficient (see Fig. 8).

As mentioned above, in case 2 ^**, to control the transmission coefficient, it is necessary not to rigidly fasten the flange of the hydraulic drive sleeve to the body, but to force control the rotation speed of the main torque converter (Fig. 8, item 2) in its entire range from ω _tr to + ω _tr that is, from braking the free rotation of the main torque converter 2, rotating in the direction opposite to the rotation of the primary shaft 5, to forced rotation of the main torque converter in the direction of rotation of the primary shaft.

 Let us try to implement this function engineeringly without changing the right-hand side of FIG. 7 - the power part, and into the left (control) part, introduce a device that allows controlling the translational movement "left - right" to carry out the function of controlling the angular speed of rotation of the main torque converter (see Fig. 8).

The right side of FIG. 8 fully repeats FIG. 7, except that the torque converter "smoothly" passes on the left side to the secondary cylinder 4, turning from the torque converter on the right side to the hydraulic turbine on the left side, and the inner part of the secondary cylinder 4 on the left side of the pattern does not have blades (smooth surface). The secondary cylinder is a hydroturbine only on the left side of the figure. Two new parts appeared: 9 - an additional primary cylinder (additional hydraulic pump). Its design is slightly different from the main primary cylinder in that the blades on it alternate with a smooth surface, 10 - an additional torque converter. The difference between the additional torque converter (pos. 10 in Fig. 8) and the torque converter (pos. 2 in Fig. 7) is that it has a constructive function of translational movement along the central axis left-right through the torque converter sleeve 1, which in this case is rigidly fixed to the body and does not have the ability to rotate around its axis by means of rods 11. The torque converter (pos. 10 of Fig. 8) as well as the torque converter (pos. 2 of Fig. 7) does not rotate, but has only translational movement along the axis of rotation to the left -right. The main torque converter (pos. 2 of Fig. 8) differs from the torque converter (pos. 2 of Fig. 7) in that it provides:
1. The possibility of translational movement along the axis of rotation;
2. Free and forced rotation in both directions.

 The additional torque converter 10 of FIG. 8 is also made sectionally like an additional primary cylinder: the region that corresponds to the smooth surface of the primary cylinder does not pass liquid through itself.

 Thus, when moving the additional torque converter to a distance h to a distance h, we completely block the exit channels of the secondary cylinder of the oil medium perturbed by the blades, thereby eliminating the inhibition of the main torque converter (Fig. 8, item 2), which, in turn, will stop rotating the secondary cylinder (Fig. 8, item 4). This situation corresponds to a neutral gear.

Continuing the translational movement of the additional torque converter either in the same direction or vice versa, we reduce the overlap, and part of the liquid perturbed by the blades of the additional primary cylinder through the rotary channels of the additional torque converter falls on the blades of the main torque converter (which acts as a secondary cylinder in region P), thereby creating braking to the proper rotation of the main torque converter, stopping it at certain values of h and rotating it to the side, etc. oppositely own rotation for h> h _{a stop} under the action of the difference of the forces applied to it by two different streams: an adjustable flow from secondary converter and force arising mainly a torque converter during a turn of the primary flow from the main cylinder (Fig 8 pos 5..).

 So, the design of the hydraulic actuator in FIG. 8 provides control of the gear ratio, control of the beginning and end of movement (neutral gear) in one direction of rotation of the shaft of the secondary cylinder. Now we need a design that provides the above functions in both directions of rotation of the secondary shaft: forward and reverse (reverse) without changing the direction of rotation of the primary shaft. To solve this issue, a sectional type construction is proposed. In FIG. 8, the direction of rotation of the shaft of the secondary cylinder is determined by the direction of rotation of the knee in the main torque converter. It is clear in which direction the “jet” strikes, the turbine rotates there, therefore, it is proposed that the main torque converter be made not with an integral structure with the blade knee turning in one direction, but sectionally, that is, in turn to put the left and right sections, alternating them one after another and combining in one package.

 At the same time, the design of the main primary cylinder (Fig. 8, item 5) must also be changed and made structurally similar to the additional primary cylinder, that is, the blades of the main primary cylinder will alternate with a smooth cylindrical surface. If you now change the type of landing of the main primary cylinder on the shaft (Fig. 4, item 14), that is, replace the rigid attachment of the cylinder to the shaft with a mount that allows the main cylinder to move left-right along the shaft, then by moving the primary primary cylinder to the leftmost position, the disturbed liquid will pass only through the left sections (the right ones will turn out to be unperturbed).

 Thus, the shaft of the secondary cylinder will rotate to the left, thereby providing a backward movement, and accordingly, the extreme right position of the primary primary cylinder will ensure the rotation of the shaft of the secondary cylinder to the right, thus ensuring forward movement. In order to maintain controllability by moving forward and backward, an additional torque converter will also have to be structurally changed (Fig. 8, item 10). It is practically sectional here, that is, the right section alternates with an empty section (smooth surface). Now we will replace the “empty section” with the left section, since the translational movement left-right on the additional torque converter is already provided for in the design of FIG. 8, replacing the empty section with the left one makes it possible to control the rotation speed of the main torque converter (Fig. 8, item 2) in two directions, forward and reverse, thus ensuring controlled movement in the forward and reverse directions, with corresponding combinations of relative positions of elements. By the way, for control convenience, it may be more convenient to move the main cylinder relative to the main torque converter, but not the main torque converter, relative to the main cylinder, as the second fulcrum of the main torque converter is in contact with the torque converter sleeve (Fig. 8 pos. 1) therefore, it will be convenient enough to pass a couple more rods through the sleeve to control the movement of the main torque converter.

 Thus, two pairs of control rods will pass through the sleeve: the first pair, as shown in FIG. 8 poses 1, controls the additional torque converter (pos. 10), moving it left-right, and the second pair of rods (it is not shown in Fig. 4, but will be perpendicular to the plane) controls the main torque converter, moving it left and right (working for this pair the positions are only two extreme positions, since these rods provide movement either forward or backward). Each pair of rods with its left ends is connected to its stepper motor (there are two of them).

 One stepper motor electrically controls the change in the gear ratio, the beginning and end of the movement depending on the relative location of the additional primary cylinder and the additional torque converter, another stepper motor controls the forward and backward movement depending on the relative position of the primary cylinder and the main torque converter. Both steps are attached as well as the sleeve to the shell body. The shell body, in turn, is rigidly attached to the engine block, has an opening for entry through the stuffing box of the primary shaft and exit through the stuffing box of the shaft of the secondary cylinder. It performs the function of protecting the secondary cylinder and collecting the squeezed out oil from the area of increased oil pressure through the shaft rods of the torque converter sleeve and through the rings of sliding bearings. To return the squeezed out oil to the working area, it is advisable to place an oil pump near the stepping motors and pump the squeezed out oil back into the working area through the torque converter bushing and pressure reducing valve. The pump is mounted on the casing, driven through the gear from the input shaft.

 Thus, inside the casing, it is necessary to place three service devices: two stepper motors and an oil pump, which is also a difficult engineering task.

 So, we considered the possibility of creating an electrically controlled device that can control the movement of the car forward, backward, changing the gear ratio, the beginning and end of the movement forward and backward, to perform these functions gently, smoothly and without friction elements, without gears. The contact of rotating parts occurs only in the bearing units (pivot points), and the main interaction of the parts occurs in the area of the blades through a "rigid" oil medium (medium under pressure). Therefore, the wear of the main assembly is determined by the wear of the bearings, oil pump and stepper motors.

 Since the control of this unit is electric, with an appropriate order, you can make an additional function - radio control. For example, upon receipt of the appropriate code on the air, the car can be stopped. Or, for example, when an appropriate signal arrives on the air, you can reduce the speed of the car, that is, in places where the speed limits, you can force the speed of passing cars equipped with this type of automatic transmission. The need for such a design is more than obvious. The cost of such a design in mass production can be about 5 times cheaper than existing automatic transmission designs that perform the same functions as this one, with the exception of one - radio control. The new design, in addition to simplicity, reliability and low cost, is also radio-controlled and, therefore, promising.

 Based on the foregoing, it is proposed to call it a hydrodynamic continuously variable transmission.

Claims (11)

 1. Hydrodynamic stepless gearbox containing primary and secondary shafts arranged coaxially and connected respectively to the engine flywheel and cardan shaft, a hydrodynamic torque converter located in a housing rigidly connected to the engine block, characterized in that in the closed volume of the hydrodynamic circulation a torque converter are located the primary cylinder of the pump and the secondary cylinder of the pump, rigidly mounted on the input shaft, the reactor cylinder the auxiliary converter torque located around one another in the primary pump primary cylinder, the primary converter reactor cylinder, located around the primary pump cylinder and around the secondary converter cylinder, the secondary cylinder turbine rigidly connected to the secondary shaft, forming an enclosed cylindrical space around the primary torque converter, torque converter, primary pump cylinder and additional primary th pump cylinder, wherein the primary and secondary torque converters are formed with respective control rods.  2. The hydrodynamic continuously variable transmission according to claim 1, characterized in that the additional oil pump with a gear drive of rotation from the input shaft is located inside the housing, the pressure reducing valve of the pump controls the pressure in the working area in dynamic mode.  3. The continuously variable hydrodynamic gearbox according to claim 1 or 2, characterized in that the rods of the main torque converter and the additional torque converter are moved using two stepper motors.  4. The continuously variable hydrodynamic gearbox according to claim 3, characterized in that the rods of the main torque converter and the additional torque converter are moved using the electronic control unit.  5. The hydrodynamic stepless transmission according to claim 4, characterized in that the rods of the main torque converter and the additional torque converter are moved using a radio-controlled signal.  6. The hydrodynamic stepless transmission according to claim 1 or 2, characterized in that the movement of the rods of the additional torque converter is carried out using a stepper motor.  7. The continuously variable hydrodynamic gearbox according to claim 6, characterized in that the movement of the additional torque converter rods is carried out using an electronic control unit.  8. The hydrodynamic stepless transmission according to claim 7, characterized in that the movement of the rods of the additional torque converter is carried out using a radio-controlled signal.  9. Hydrodynamic continuously variable transmission according to any one of paragraphs. 3 and 7, characterized in that the anti-theft system, as an integral part of the vehicle structure, is inscribed in the design of the electronic control unit.  10. The hydrodynamic continuously variable transmission according to claim 5 or 8, characterized in that it is configured to control the vehicle speed to the level of the maximum allowable speed by radio.  11. Hydrodynamic continuously variable transmission according to claim 5 or 8, characterized in that it is configured to force radio control of the vehicle speed to stop and subsequently detect a stolen vehicle.

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