RU2013629C1 - Engine - Google Patents

Abstract

FIELD: automotive engineering. SUBSTANCE: engine has compressor with inlet, external combustion chamber connected with compressor through first by-pass having fuel spray nozzle and ignition source disengageable during operation, expander with inlet communicating with combustion chamber through second by-pass, and working members located in compressor and expander and operatively connected to output shaft. Compressor, combustion chamber, and expander have separate cooling systems, whereas combustion chamber and expander have a volume exceeding the volume of the compressor. Operative connection of working members with output shaft is fashioned as a planetary mechanism with two bevel pinions, two satellite gears, and carrier with toothing. The working member of the compressor is connected to one of the bevel pinions, output shaft is connected to another bevel pinion, and working member of the expander is connected to the toothing of the carrier. Each working member can be fashioned as at least one pair of blades. EFFECT: improved design, enhanced efficiency. 11 dwg

Description

 The invention relates to the field of heat engines, in particular to heat engines, and can be used as a power plant of any vehicle or stationary device.

 Among the technical solutions of heat engines, two main categories are known: engines in which the products of combustion of air-fuel mixtures play a dual role, being both carriers of thermal energy and a working fluid (conventional ICE), and an engine in which the role of products of combustion of air-fuel mixtures is reduced to the role of sources of heat energy supplied from the outside and used to heat a gaseous working fluid in a closed system, i.e. engines operating according to the Stirling and Willumier cycles (external combustion engines and I). As a rule, helium or hydrogen are used as a working fluid.

 Despite the intensive development and improvement of the ICE design, the result of which is an increase in their output parameters in various aspects (liter power and torque, economy, reduction in the concentration of harmful emissions, durability, etc.), their schematic diagram remains practically unchanged, and improvement is carried out as a rule, by indirect means and methods (electronic ignition systems, fuel injection, exhaust gas converters, etc.), or quantitative improvements are not applied fundamentally character (increase in flow section valves regulating flow areas of the intake and exhaust tracts, increase of the compression ratio, and others.). The exception is rotary piston engines (RPD), although they are only a rotary piston version of conventional internal combustion engines. Most of the shortcomings of the internal combustion engine is a consequence of the imperfection of the basic scheme of the internal combustion engine, which consists in the fact that its entire working cycle takes place in one working space (cylinder or chamber, in the case of RPD) and is distributed in time with a sequential alternation of clock cycles.

 The main disadvantages of this scheme include: limitations on the flammability of the mixture with respect to α (excess air coefficient), and therefore the need for artificial re-enrichment of the mixture (mainly ICE with spark ignition); incomplete combustion of the mixture due to the transience of the combustion-expansion process; restrictions on the types of fuel used by the octane or cetane number, as well as types of fuel; irrational use of the working space, exclusion from it of the volume of the combustion chamber that is not involved in the expansion process (constant volume of the chamber); incomplete use of energy of combustion products due to insufficient expansion (high residual pressure at the beginning of the exhaust and the presence of residual gases in the combustion chamber; high concentration of harmful components in the combustion products, the inability to maintain optimal temperature for intake and expansion of the temperature regime and, as a result, poor filling cylinder (chamber) at the inlet and insufficient use of energy of the combustion products during expansion.

 The above list of shortcomings of the known ICEs is caused only by the lack of distribution of the working cycle in space and the combination of the space of the combustion chamber in the working space of the engine cylinder. This problem finds a partial specific solution by completely isolating the working space from the combustion chamber in the Stirling engine, the principle of which is based on alternately heating and cooling the working fluid using a regenerator that selects and returns heat to the working fluid, i.e., by alternately creating the so-called. hot and cold spaces. The heating of the hot space is carried out by the heat generated during the combustion of fuel in the burner, and the cooling of the cold space by means of coolers. The energy (pressure) generated as a result of the expansion of the working fluid when heated acts on the power piston and is transferred to the output shaft through the rhombic drive.

 The advantage of the Stirling engine is that due to the complete separation of the heat source from the workspace, the following is achieved: full use of the workspace for the work process (cycle); unpretentiousness to fuel (multi-fuel) and the fundamental possibility of working on any heat source - from liquid fuel to atomic and solar energy; the absence or low concentration of harmful emissions due to stability and complete combustion of fuel; silent operation and the practical absence of vibration; high fuel efficiency; a more favorable (compared with ICE) torque characteristic, which allows to significantly simplify the gearbox (reduce the number of stages and gear ratios), etc.

 However, with obvious advantages, the Stirling engine has significant disadvantages. The main one is the design complexity with the presence of a large number of auxiliary mechanisms, which include a rhombic drive, a displacer, a regenerator with coolers, a heating piping system that provides heat transfer from the combustion chamber to the burner to the working fluid, etc.

 The second significant drawback of the Stirling engine is the inertia of the cooling and heating system of the working fluid, i.e., the limited rate of transition of the working fluid from a cold to a hot state, and vice versa, which has a negative effect on the engine output parameters (power, output shaft speed) and requires appropriate characteristics of the working fluid, primarily in terms of heat resistance. This explains the fact that for almost 180 years of the idea's existence, Stirling engines are not widely used, and their production is piece-wise and practically does not go beyond the scope of experimental or prototypes.

 Known attempts to distribute the duty cycle in space and in relation to the ICE scheme. These include, for example, the invention of ed. St. N 80445, in which the working process is spaced into three cylinders, one of which is enriched or rich air-fuel mixture, and the other two are air. After ignition of the mixture, the valves of the bypass channels connecting the first cylinder with the other two open in the first cylinder. Since the pressure of the air compressed in these cylinders to the beginning of the combustion-expansion process exceeds the pressure of the mixture burning in the first cylinder, air flows into the first cylinder while fuel is injected through nozzles installed in the bypass channels. After equalization of pressure, the reverse process occurs, i.e., the burning mixture is displaced from the first cylinder into the other two, which are thus included in the combustion-expansion process. With this scheme, the engine is running on lean mixtures when igniting an enriched or rich mixture in the first cylinder and partially separating the cycle in space. In addition, this can significantly increase the overall degree of compression, and therefore expansion, which generally increases the thermal efficiency of the engine, and also gives some other advantages.

In an improved version by ed. St. N 128231 provides only two cylinders (or two groups of cylinders), in one of which a rich mixture is supplied, and in the second - only air. Both cylinders are connected by a connecting (bypass) channel without valves. In the air cylinder, the volume of the combustion chamber is zero, therefore, all air at the end of compression is displaced into the main cylinder; moreover, the processes in the air cylinder are delayed by 21-22 ^about (in the angles of rotation of the crankshaft) in relation to the processes in the main cylinder.

 This option, in comparison with the previous one, has a simpler technical solution and has some additional advantages.

 However, with all the advantages, both options have significant disadvantages. Firstly, in each of them one can see an attempt to implement the well-known idea of a prechamber engine and / or the idea of an engine with layer-by-layer ignition of a mixture, but in a complicated form. Secondly, a partial separation of the cycle in space (along two or three cylinders) does not exclude the occurrence of processes (inlet, compression, combustion-expansion) in each of the cylinders, which leads to a decrease in the filling coefficient. Thirdly, the combustion of the air-fuel mixture in the working space, and not in the external combustion chamber, inevitably leads to a decrease in the parameters (temperature, pressure) of the working fluid, and therefore, the output parameters of the engine, including its thermal efficiency. In four, the flow of the working cycle in the working space entails underutilization of the thermal energy of the working fluid, a significant part of which is diverted to the cooling system and exhaust system. Fifthly, such a scheme excludes the possibility of regulating (limiting) the supply of the quantity of the working fluid to the working space, which also leads to insufficient use of the thermal energy of the combustion products due to incomplete expansion during the working stroke.

 Also known is an engine comprising a compensator with an inlet element, an external combustion chamber connected to the compressor through the first bypass element and having a fuel nozzle and an ignition source that can be switched off during operation, an expander with an exhaust element in communication with the combustion chamber through the second bypass element and working bodies located in the compressor and expander and kinematically connected with the output shaft, moreover, the compressor, combustion chamber and expander are individually provided E cooling system and the combustor and expander configured to volume exceeding the volume of the compressor.

 The aim of the invention is to eliminate the main disadvantages of analogues and prototypes and obtain additional benefits: increasing fuel efficiency; reduction in the concentration of harmful components in combustion products; reduction of vibration and noise; maintaining optimal thermal conditions; optimization of torque characteristics; increase in specific power; the possibility of conversion into an external combustion engine.

 Improving fuel efficiency is achieved in several ways that are used comprehensively. The first way is to create the possibility of the engine working on lean fuel-air mixtures, as well as on various grades and types of fuels, including low-octane gasolines, diesel fuel, other types of liquid fuels, including monoesters of fatty acids, vegetable oils, shale and coal fuels origin, coal powder mixed with liquid fuel (suspension), fuel of wide fraction composition, various types of gaseous fuels, including hydrogen, ammonia, etc. The second way is to provide conditions for the maximum possible the full combustion of the mixtures. The third way is to rationalize the use of thermal energy of the combustion products of the mixture. Technically, all three paths in the complex come down to the distribution of the entire process not in time but in space, i.e., to dividing the working space of the engine into three parts - compressor (for intake and compression cycles), working (for expansion and output cycles) and a combustion chamber carried out beyond the limits of said parts, connected to them by bypass pipelines.

 The removal of the combustion chamber outside the working space eliminates the presence of moving parts in it, and therefore the need for lubrication and cooling, which, in turn, allows it to maintain a maximum temperature, limited only by the heat resistance of the material of its walls and, as a result, self-ignition of air-fuel mixtures a wide range of α and grades and types of fuel. In addition, a sufficiently large volume of the combustion chamber, limited only by design considerations, allows several charges of the air-fuel mixture to be in it at the same time, which proportionally increases the available combustion time of the mixture, i.e., in combination with high temperature increases the completeness of its combustion.

 The separation of the combustion chamber from the working space of the engine makes it possible to dose the supply of the working fluid, which is the product of the complete (or close to it) combustion of the air-fuel mixture, into the working part, and therefore to optimize (maximize) the utilization of thermal energy of the working fluid during the expansion process.

 The solution to this problem allows us to implement the initial idea of a ceramic adiabatic engine, which for known reasons in its pure form did not give the expected results. A similar problem is solved by implementing the idea of the Stirling engine. However, unlike him, in this case, the working fluid is the direct products of combustion of the air-fuel mixture, the combustion chamber is connected by a pipeline to the working part and the compressor part, which supplies the air-fuel mixture to the combustion chamber, etc.

 The separation of the working space of the engine into the compressor and working parts eliminates the direct dependence of the degree of compression on the degree of expansion (as the reciprocal of the degree of compression), which makes it possible to more deeply expand, and therefore, more fully utilize the thermal energy of the working fluid. This is achieved either by a corresponding increase in the piston stroke of the working part (in the piston version), or by an increase in the difference in the angles of rotation of the corresponding blades (in the case of RPD), or by an increase in the volume of the working part relative to the dosed volume of the working fluid entering the working part, or a combination of these measures. Theoretically, this makes it possible to bring the utilization of thermal energy of the combustion products to a “cold exhaust”, that is, to zero or close to the pressure of the start of release.

Reducing the concentration of harmful components in the combustion products is carried out automatically by solving the first of these problems, at least in terms of CO and CH, as well as lead content, since CO and CH are products of incomplete combustion of air-fuel mixtures, and lead is a consequence of the use of thermal power plants to increase octane numbers (ethylation) of gasoline. With sufficient completeness of combustion of air-fuel mixtures, the problem of converting CO to CO ₂ , as well as unburned hydrocarbons, is largely eliminated, and without the use of special devices such as afterburners, thermal and catalytic converters, and other devices that complicate the design, requiring large additional costs that reduce engine power and increasing fuel consumption. For the normal operation of these devices, the presence of combustible components in the exhaust gas is required. This leads to the fact that a significant part of the fuel that must be burned in the engine burns out in the exhaust tract, as for lead, the engine's unpretentiousness to fuel completely eliminates the need to use TPP additives in the fuel, and therefore the lead content in the combustion products. In relation to such toxic components as NO _x , S, etc., experimental verification is required, however, it can be assumed that their concentration will be no more than that of prototypes and analogues. In particular, the concentration of toxic components in the combustion products of the Stirling engine for a standard urban driving cycle is SO-0.007-0.03 wt. %, CH-100-200 ppm. Achieving this level in relation to the proposed technical solution seems to be quite possible.

 Reducing the level of vibration and noise is reduced to eliminating the causes that cause them, or to suppressing the intensity of the actual vibration and noise.

 The main sources of vibration and engine noise are the presence of a crank mechanism, a change in pressure during the cycle and when loads change, mechanical knocks and noises arising in various mechanisms, piston impacts on cylinder walls, detonation, noise in the exhaust system.

 In the Stirling engine, this problem is solved by means of a rhombic drive that eliminates lateral loads on the pistons, equalizes the pressure throughout the cycle, the lack of a gas distribution mechanism, the absence of detonation, stable and constant combustion of the air-fuel mixture.

 In the invention, the problem of reducing vibration and noise is solved by moving the combustion chamber out of the working space and isolating the compressor part from the total working space of the engine, which ensures smoothing of pressure fluctuations upon ignition of the next portion of the air-fuel mixture due to the almost continuous combustion process due to the portioned supply of the mixture and the damping effect of the products combustion in the combustion chamber. The pressure pulsation in the combustion chamber arising from such a process is inversely related to the ratio of the volume of the combustion chamber to the volume of the compressor part, and, as this ratio selected during design increases, it asymptotically approaches zero, regardless of the absolute value of the working pressure in the chamber combustion. In practice, the magnitude of this ratio is limited only by design considerations (overall dimensions) and can be brought up to 3-5. For a conventional ICE with a compression ratio of about 9, it is about 0.1. This means that the pulsation of the operating pressure in the engine can be reduced, in comparison with a conventional ICE, by 30-50 times, i.e., it is almost reduced to zero.

 The solution to the noise problem due to detonation combustion of air-fuel mixtures is achieved in the same way, that is, by moving the combustion chamber out of the workspace and increasing its volume relatively.

 The problem of noise caused by piston impacts on the cylinder walls, the gas distribution mechanism, the crank mechanism, etc., in the piston version of the engine is only partially solved due to the stability of the operating pressure in the steady state and the smoothness of its change in transition modes. In the RPD variant, this problem does not arise due to the lack of the mentioned mechanisms in the traditional design.

 The problem of noise in the exhaust tract, due to a sufficiently high pressure at the beginning of the exhaust, is solved automatically when solving the above problems of fuel economy and toxicity of combustion products, i.e., by continued expansion of the working fluid and, as a result, a decrease in pressure and temperature of the exhaust, as well as dosing of the feed rate the working fluid into the working cylinder or chamber, which also allows for continued expansion, but in an indirect way.

 In general, the solution to the problem of vibration and noise is achieved by methods similar in nature to the methods used in an external combustion engine, however, the main distinguishing feature in this case is the simplicity of the technical solution, eliminating the use of a number of additional mechanisms.

 To optimize the output parameters (power, torque, fuel efficiency, etc.) of a heat engine using the combustion products of air-fuel mixtures as a working fluid, it is necessary to optimize its temperature regime. This optimization consists in the selection of special temperature conditions for elements or groups of elements (cycles) of the entire working cycle. For example, for inlet and compression, it is desirable to maintain a low temperature, which allows for sufficient filling and to avoid unnecessary energy (power) loss in compressing the mixture, for the combustion-expansion process, high temperature in order to more fully utilize the thermal energy of the working fluid. Attempts to solve this problem are being undertaken by many foreign firms. In particular, to increase the filling coefficient, charging was widely used, which entails the use of turbochargers or mechanical superchargers, and, as a rule, their use is accompanied by the installation of charge air cooling systems. To increase the temperature and the degree of heat recovery, attempts are being made to implement the idea of an adiabatic engine. However, all these methods and tools give only a partial effect and, with a significant complication of the design, do not solve the problem as a whole, and sometimes lead to a dead end.

 The invention allows to solve this problem as a whole, with great effect and without undue complexity of the design. This is achieved by distributing the working cycle into space and dividing it into three elements or processes - intake and compression, including bypass, combustion, expansion and exhaust, which allows you to maintain optimal temperature conditions for each element or each group of cycle elements through separate cooling systems or the corresponding distribution of coolant flows. Thus, the high cooling rate of the compressor part allows it to maintain a low inlet and compression temperature, less intensive cooling of the working part — the high temperature of the expansion process with a natural decrease in this temperature by the end of the expansion stroke and the beginning of the exhaust stroke, as for the separate combustion process, cooling removed from the working space of the combustion chamber may either be completely absent or minimized due to the heat resistance of the material.

 As an additional effect, this solution allows you to maintain a relatively stable temperature of the compressor and working parts, and therefore to eliminate changes in the value of temperature gaps, and, accordingly, reduce friction losses.

 For existing internal and external combustion engines, the torque value within certain limits, at least within the operating mode, varies in direct proportion to the shaft speed, which entails the need to use additional devices such as gearboxes. This is one of the fundamental differences from electric motors and steam engines. Ideally, it is desirable to have an engine whose torque characteristic would have an inverse relationship. This problem is partially solved in the Stirling engine. However, the Stirling engine changes only the degree of this dependence, and not the fundamental nature of it. The invention allows to solve this problem in principle, that is, to change the very nature of the dependence of torque on the frequency of rotation of the output shaft, or to obtain its "inverted" characteristic. The presence of an inverse dependence of the moment on the output shaft rotation speed means that the torque reaches its maximum value at a speed of the engine output shaft equal to zero (fixed shaft), and the minimum - at the maximum rotation speed of the latter. In practice, this eliminates the need for a gearbox and clutch and gives the following additional advantages: automatic adaptation of the engine torque value, for example, to the vehicle’s driving conditions (resistance to movement) or the momentum required to drive some mechanism, regardless of load (feed rate mixtures); engine operation in optimal mode, i.e. in maximum torque mode at full load (according to an external characteristic) or any partial load; stable operation of the engine without spontaneous stop, regardless of the required torque on the output shaft; smooth change in the speed of the output shaft during acceleration or deceleration; reduction in specific fuel consumption due to engine self-tuning to the optimal operating mode, i.e., the maximum torque mode for a given load, and therefore the maximum utilization of the thermal energy of the working fluid.

 For well-known internal combustion engines and external combustion engines, an increase in power is achieved (ceteris paribus) either by increasing the working volume of the cylinder, or by increasing the number of cylinders (sections, in the case of RPD), or a combination of both methods. At the same time, the task of increasing the uniformity of engine operation, i.e., reducing the level of vibration, is being solved. However, these measures, both individually and in combination, give the corresponding effect only in relation to absolute, and not specific (relative engine power, i.e., quantitative result) and do not affect the solution of the problem in a qualitative aspect.

 In the piston version, the solution to the problem of increasing power is achieved mainly by increasing the working pressure in the combustion chamber, and, consequently, by increasing the torque due to the use of a planetary mechanism that provides the corresponding value and intensity of the supply of the air-fuel mixture to the combustion chamber, i.e., the number of charges of the mixture per working stroke (expansion). This allows you to increase power without changing the working volume of the compressor and working parts of the engine, i.e., increasing the power by units of the total working volume of the engine.

However, a more radical method of solving this problem in the complex is to increase the number of pairs of blades in the compressor and working chambers of the RPD variant. Since in this embodiment the working volume of the chamber is the difference between the volume of the inner space of the chamber and the volume occupied by the blades and shafts, moreover, it does not depend on the number of blades and remains constant, it becomes possible to increase the number of pairs of blades with a constant working volume. This gives a multiple reduction in duration and, consequently, an increase in the number of cycles per revolution (360 ^° ) of the chamber shaft, both compressor and working, and, therefore, by one revolution of the output shaft. An increase in the number of cycles with a constant working volume of the camera is equivalent to a corresponding increase in the number of cameras. For example, an engine with one pair of blades is equivalent in its characteristics to a single-chamber (single-section) RPD, with two - two-section, three - three-section, etc. The number of pairs of blades is limited only by design considerations, and in this regard, the likely limit of their number is 4 pairs of blades or 8 blades. Within the range of 1-4 pairs of blades, the cycle duration varies in multiples from 360 ^о to 90 ^о , and the number of cycles is from 1 to 4. This solution allows you to: increase the specific (liter) engine power in comparison with the rotor-piston analogue having 2 pairs of blades but working in a four-stroke internal combustion engine mode; reduce the intensity (level) of vibration compared to the same option.

 Obtaining these results is achieved only by increasing the number of pairs of blades without taking into account the receipt of the corresponding results by other methods.

 Such a technical solution is one of the main distinguishing features of the invention.

 However, one should take into account the specific feature of the proposed engine. Since the power value of any engine is described by a product, one of the factors of which is the shaft rotation frequency, it would be logical to assume that for this factor zero, the product, i.e., the power value, is also zero. However, the presence of a planetary mechanism allows the engine to work even with a stationary output shaft and when rotating only the shafts of the compressor and working parts. Hence, the power developed by the engine is described by the product, the factors of which are the moment and speed, which apply only to the shaft of the working part of the engine, and all this power is spent only on the drive of the compressor part shaft, that is, for the engine to work "on itself". When the torque reaches a value sufficient to drive the output shaft, the power on the output shaft acquires a value other than zero, and, ceteris paribus, varies in functional dependence on the speed of the output shaft. In this case, the power distribution to the drive shafts of the compressor and working parts is proportional to the distribution of torque between them, determined by the internal gear ratio of the planetary mechanism, set during design depending on the purpose of the engine. If the task is to increase power by reducing the speed of the output shaft, the gear ratio is selected, which allows to increase the speed of the compressor relative to the output shaft, and vice versa.

 All the main listed distinctive features of the invention relate to a variant of a semi-external combustion engine. However, this option can be converted into an external combustion engine without significant alterations. For this, it is only necessary to provide heat from the outside to the walls of the combustion chamber, turning it, thus, into a heating chamber for the working fluid. In this case, as in the case of the Stirling engine, any source of thermal energy (liquid fuel, coal and firewood, the Sun, etc.) can be used. Atmospheric air (in the case of an open system) or any gaseous body can be used as a working fluid. In both cases, there is no need for sequential alternation of heating and cooling cycles of the working fluid, and these processes themselves become continuous, in contrast to the Stirling cycle.

 In the case of using air taken from outside, there is no need for its special cooling, since the temperature of the atmosphere is much lower than the temperature of the air in the heating chamber. The regulation of the output parameters of the engine is carried out in this embodiment either by controlling the intensity of heating of the working fluid by controlling the flow (supply) of fuel, or by throttling the flow of the working fluid at the inlet to the compressor, or by a combination of both methods. In principle, it is possible to use for this purpose and control (throttling) the flow of hot working fluid from the heating chamber.

 Thus, with some slight change in the concept of the proposed engine, which consists in supplementing its design with a burner or other heating device installed outside the combustion chamber, supplying heat to the combustion chamber and converting it into a heating chamber such as a steam boiler, it is converted from a semi-external combustion engine into an external combustion engine without a number of additional devices. At the same time, some additional disadvantages are excluded, in particular, the cyclic heating and cooling of the working fluid, the influence of its thermal inertia, as well as the system as a whole, on the output parameters, the inability to use an open system, the complexity of the circuit diagram, and therefore the technical solution (elimination of a number of mechanisms and systems).

The invention is presented in several versions and sub-options, which differ from each other in technical solutions, execution and, as a result, output parameters. These include:
Option I - a semi-external combustion engine with translational movement of the pistons and a common crankshaft for the compressor and working cylinder;
Option II - a semi-external combustion engine with translational movement of the pistons and separate crankshafts of the compressor and the working cylinder, kinematically connected through two elements of the planetary mechanism (differential), the third element of which is directly connected to the output shaft.

 Option III - a semi-external combustion engine in a rotary piston design (RPD), which is a version of option 1.

 Option IV - a semi-external combustion engine, which is a version (sub-option) of option II in the performance of the RPD.

 Option V - an external combustion engine, which is a version (sub-option) of option I.

 Option VI - an external combustion engine, which is a sub-option of option II.

 Option VII - an external combustion engine, which is a sub-option of option III.

 Option VIII - an external combustion engine, which is a sub-option of option IV.

 The essential features of the invention include: the separation of the working space into three parts - the compressor, the working part (cylinder, chamber) and the combustion chamber with the distribution of the elements of the duty cycle, i.e., intake and compression-bypass processes (compressor), combustion process ( combustion chamber), processes of bypass-expansion and exhaust (working part) with the simultaneous flow of all processes within one cycle, which allows you to maintain the optimum temperature in each part and, as a result, the exclusion or reduction to mini the mind of burning the mixture during expansion, increasing the filling coefficient at the inlet, self-ignition when the engine is heated, the air-fuel mixture of various compositions both in α and in types and grades of fuel, the completeness of combustion of the mixture due to the high temperature in the combustion chamber and the increased duration (available time) of combustion mixtures and, as a result, increased fuel economy and environmental friendliness, as well as the lack of need for a mechanism for regulating the moment of ignition and the need for a system ignited Iya generally when the engine is warm; the use of separate cooling systems or different cooling intensities for the compressor and working parts in order to optimize the temperature regime of each of them with the highest possible heat resistance of the insulation material of the combustion chamber (adiabatization effect); the continuity of the combustion process of the mixture as a result of feeding another portion of it for each cycle and the presence of several charges of the mixture in the combustion chamber (depending on the volume of the chamber), which ensures a constant working pressure in the steady state, smooth change in transient conditions and, as a result , a significant reduction in the level of vibration and an increase in the uniformity of engine operation, as well as detonation-free combustion of each subsequent portion of the mixture due to the small relative magnitude of the pressure increment; the removal of the combustion chamber beyond the working part (cylinder, chamber) in combination with the appropriate selection of the gas distribution phases allows metering the flow rate of the working fluid, which increases the degree of utilization of its thermal energy during expansion and, as a result, fuel efficiency and thermal efficiency; separation of the common shaft of the compressor, the working cylinder (chamber) and the output shaft into three corresponding parts and the establishment of a kinematic connection between them through a planetary mechanism (differential), which ensures self-regulation of the torque on the output shaft in inverse proportion to the speed of the latter in the range from zero to maximum and from maximum to zero while ensuring optimal engine operation, that is, maximum mode for these specific torque conditions; the ability to select the appropriate internal gear ratio of the planetary gear (differential) and / or use a compressor and a working cylinder of various volumes allows you to change the limits of the range of values of the torque conversion coefficient in relation to the specific purpose of the engine; the separation of the working space into four parts — the compressor, the working part (cylinder, chamber), the heating chamber and the combustion chamber with complete isolation (except thermal) from the engine’s working space and the supply of heat from it to the heating chamber (in a variant of the external combustion engine), which It allows continuous heating of the working fluid, eliminating the influence of thermal inertia of the working fluid on the engine output parameters, using atmospheric air as the working fluid, and simplifying the design of the engine by by excluding from it all devices and systems for alternating (cyclic) heating and cooling of the working fluid, i.e., to ensure the transformation of the semi-external combustion engine into an external combustion engine with a significant simplification of the design and an increase in its output parameters and an additional reduction in the level of vibration; the possibility of increasing the pairs of blades (blades) in the case of a rotary-piston version to a value limited only by structural and strength considerations (almost 4 pairs), which allows for a constant chamber volume to increase the number of cycles by one revolution of the shaft and, as a result, proportionally increase specific power and reduce vibration, i.e., get the effect equivalent to the effect of increasing the number of cylinders (chambers); the presence of a compressor allows for the forced supply of a cold working fluid (air) to the heating chamber (in the variant of an external combustion engine), which contributes to an increase in gas exchange intensity and, as a result, an increase in the engine output parameters (power, torque).

 The listed essential features are complex with a complex effect and all relate to distinctive features of the prototype (prototypes). Of these, the second symptom is sufficient in all cases to which the requested amount of legal protection applies.

 The invention is illustrated by drawings, where in FIG. 1 shows a schematic diagram of an engine in embodiment 1; in FIG. 2 is a calculated comparative PV diagram; in FIG. 3 is a schematic diagram of option II; in FIG. 4 - the general nature of the dependence of the torque on the output shaft of the engine on the frequency of its rotation; in FIG. 5 is a diagram of a planetary mechanism; in FIG. 6 is a schematic diagram of an engine in embodiment III; in FIG. 7a - technical solution of the RPD chamber on the example of compressor 2; in FIG. 7b is a schematic diagram of a compressor; in FIG. 7c is a schematic diagram of a working chamber; in FIG. 8 - technical solution (8a) and circuit diagram (8b) of the regulatory mechanism; in FIG. 9 is a graph of the change in the angular velocity of the chamber shafts and the duration of the cycles for various amounts of pairs of blades; in FIG. 10 is a schematic diagram of an engine and 6 positions of its working process; in FIG. 11 - four sub-variants of the external combustion engine (in Fig. 11a - options V and VI with an open system; in Fig. 11b - options V and VI with a closed system; in Fig. 11c - options VII and VIII with an open system; 11 g - variants VII and VIII with a closed system In version I, the engine comprises a compressor cylinder (compressor) 1 with a piston 2 connected through a connecting rod 3 with a crank of a crankshaft 4, and a working cylinder 5 with a piston 6 connected through a connecting rod 7 with another crank shaft 4. Compressor 1 and the working cylinder 5 are connected bypass pipes 8 and 9 to the remote affairs of the working space of the engine with a combustion chamber 10, equipped with a fuel nozzle 11 and a glow plug or with a pulsating supply of an ignition spark 12. Bypass pipelines 8 and 9 are equipped with bypass valves 13 and 14. The compressor 1 through the inlet pipe 15, equipped with an inlet valve 16, is connected to the air cleaner or a carburetor (not shown in the diagram), and the working cylinder 5 through the exhaust pipe 17, equipped with an exhaust valve 18, with an exhaust path (not shown in the diagram). A flywheel 19 is installed on the crankshaft 4 of the engine, used, in particular, as an element of the clutch plate and to drive the shaft 4 of the engine when starting the latter.

In FIG. 2 shows the calculated comparative PV-diagram, giving a comprehensive assessment of the proposed engine in comparison with the reference (classic) ICE. All designations related to the reference internal combustion engine are marked with the index o. The abscissa axis shows the values of the total cylinder volume of both engines as a percentage formed when the pistons move from BDC to TDC and vice versa, and since the piston strokes (S and S _о ) of the compared variants differ in magnitude, each of them has its own abscissa axis. The ordinate shows the pressure P in kgf / cm ² .

Point P _a = P _ao - pressure of the end of the outlet - the beginning of the inlet. Point P _co - compression pressure at the moment of ignition of the mixture of the standard internal combustion engine, point P _with - compression pressure at the time the start of the bypass for the proposed engine. Points P _Z ; (P _Zo ), P _Z ^I (P _Zo ^I ) are the maximum pressure points during combustion of the mixture for each of the comparative engines, respectively, with the point P corresponding to the pressure of the end of the bypass and the beginning of the expansion process for the proposed one, and P _Zo to the beginning of the expansion process combustion of the mixture in the reference engine. Point P _in corresponds to the pressure of the beginning of release in the reference internal combustion engine. Point P _Z ^II is equivalent to point P _Z ^I with a corresponding increase in volume (segment P _Z ^I -P _Z ^II ) due to an increase in the dose of the working fluid coming from the combustion chamber to the working cylinder.

, где d₂₀ - диаметр конической шестерни 20, а d₂₂ - шестерни 22. Поскольку d₂₀ = d₂₂, i_o = 1.The embodiment of FIG. 3, additionally includes a planetary mechanism (differential) 19 with a bevel gear 20 of the crankshaft 21 of the compressor 1, a bevel gear 22 of the output shaft 23, a carrier (differential box) 24 and satellites 25 mounted in the bearings of the carrier 24 and connecting gears 20 and 22. The carrier 24 is connected via a ring gear 26 to a cylindrical gear 27 of the crankshaft 28, the crank of which is pivotally connected to the connecting rod 7 of the working cylinder 5. In the embodiment of the planetary mechanism (differential) 19 shown in FIG. 3, the internal gear ratio i ₀ =

where d ₂₀ is the diameter of the bevel gear 20 and d ₂₂ is the gear 22. Since d ₂₀ = d ₂₂ , i _o = 1.

In FIG. Figure 4 shows the general nature of the dependence of the torque on the output shaft of the engine on its rotation frequency, M = f (n), for various values of the internal gear ratio of the differential. The value of M is given in percent, and for the base (100%) of the adopted value of M at n = 0 and i _about = 0.5. The values of n are plotted on the abscissa of the graph, and the values of M are plotted on the ordinate. The dependence is represented by a family of five curves for i _o corresponding to i _o = 1.2; 1.0; 0.7; 0.5; 0.3.

In FIG. 5 is a diagram of a planetary gear (differential) having an internal gear ratio i _о ≠ 1. In this figure, 20 is a bevel gear of the shaft 21 of the compressor 1, 22 is a bevel gear of the output shaft 23, 24 is a carrier, 25 is a gear of the bevel gear 22; 25 - bevel gear satellites 20.

 In FIG. 6 positions have their own numbering. The engine includes a common cylindrical body, divided into 6 parts by internal partitions (walls) - compartment 1 of the control mechanism 7 of the compressor chamber 2, the combustion chamber 3, the working chamber 4, compartment 5 of the control mechanism 10 of the working chamber 4 and the compartment 6 of the planetary gear 8. Compartments 1 and 6 are closed on the outside by end caps. The regulating mechanism 7 is connected to the sun gear 17 of the planetary mechanism 8 by a shaft 9, and the regulating mechanism 10 is connected to a hollow shaft 11 with a planet carrier 26, the crown 27 of which is connected to the output shaft 12 (other combinations are possible). The compressor chamber 2 has a peripheral inlet window made in the corresponding part of the wall of the cylindrical body, with an inlet pipe 13 equipped with an air (or throttle) damper, as well as a bypass window (windows) made in the wall (partition) separating the compressor chamber 2 and the chamber combustion 3. In the wall separating the combustion chamber 3 and the working chamber 4, a bypass window (s) is made, and in the wall part of the cylindrical body, a peripheral exhaust window with an exhaust pipe 14. In the cylinder part of the wall A fuel injector 15 and a spark plug 16, either a glow plug or a spark pulsed spark plug, are installed in the housing associated with the combustion chamber 3. The sun gear 17 of the planetary gear 8 is fixedly mounted on the rear end of the shaft 9, the front end of which is connected to the control mechanism 7 at the entrance to it. The output of the control mechanism is represented by two coaxial shafts 18 and 19 with blades 21 and 20 fixed to them, respectively. Each of the shafts 18 and 19 with the blades 21 and 20 is a half-rotor, and both half-rotor assemblies form the rotor of the compressor 2. Similarly, the rotor of the working part consists of blades 22 and 23 and coaxial shafts 24 and 25, interconnected via a regulating mechanism 10 of the working chamber 4, having one common shaft 11 with the carrier 26 of the planetary gear rigidly mounted on it 8. The compressor chamber 2 and the working chamber 4 have separate cooling systems, which can be performed as per the system of the general cooling system, but with different heat intensities oobmena.

 In FIG. 7 presents: the technical solution (7a) of the RPD chamber using the example of compressor chamber 2, the circuit diagram (7b) of the compressor 2 and the circuit diagram (7c) of the working chamber 4 (Fig. 6).

 In FIG. 7a, the designations of all positions refer only to the details of the compressor 2, however, in the text in parentheses are given the designations of identical parts of the working chamber 4.

 As shown in FIG. 7a, three coaxial shafts 9, 18 and 19 (9, 22 and 23) with blades 20 and 21 (22 and 23) cantilevered on the shafts 18 and 19 (24 and 25) are placed in a cylindrical housing 2 (4). The shaft 19 (24) is made with a ledge (thickened and thinned parts), and the end face of the thickened part is located approximately in the middle of the length of the chamber. A hollow (tubular) shaft 18 (25) is put on the thinned part of the shaft 19 (24), exactly equal to the outer diameter of the thickened part of the shaft 19. The shaft 18 (25) is joined with the end face of the shoulder of the shaft 19 (24), and the joint is ground and equipped with a seal. The joint should be gapless. On the thickened part of the shaft 19 (25), the blade 20 (23) is attached, and the shaft 18 (24) - the blade 21 (22). The blades have the shape of a sector with an outer diameter equal to the inner diameter of the chamber body 2 (4) and the inner diameter (at the base) - the diameter of the shaft 18 (24) and, accordingly, the thickened part of the shaft 19 (25). On the periphery, the blades are equipped with lamellar L-seals installed in the grooves 34. The shafts 18 and 19 (24 and 25) assembled with the blades 20 and 21 (22 and 23) and with the shaft 9 are mounted on bearings 30 and 31 in the end caps 28 and 29 cameras along its axis with the possibility of rotation relative to each other. An inlet window 35 is made in the cylindrical housing of the compressor 2, and an overflow window 36 is formed in the end cap 29. In the case of the working chamber 4, an exhaust window 38 is formed in the cylindrical body thereof, and the bypass window 37 in the end cap separating it from the combustion chamber 3. The window sizes are determined depending on the choice of valve timing.

 The blades 20 and 21 (22 and 23) have internal cavities, in particular, drilled through technological holes with plugs 32 and 33. The internal cavity of the blades is filled with oil coming from the oil pump (not shown in Fig. 7) through the internal channel of the shaft 9 and used both for cooling the blades, and for lubricating the working surfaces of the blades and the associated surfaces of the housing and shafts. (Oil circulation in FIG. 7 not shown), reference numbers in FIG. 7a and 7b are given separately in accordance with the positions of the compressor 2 and the working chamber 4 (Fig. 6).

In FIG. Figure 8 shows the technical solution (8a) and the circuit diagram (8b) of the regulating mechanism as applied, as in the previous case, to the regulating mechanism 7 of compressor 2 (Fig. 6). The mechanism includes front outputs (ends) of the shafts 9, 18 and 19, mounted coaxially with the shaft 9 inside and the shaft 19 outside. A gear 28 is installed at the end of the shaft 9, which is meshed with gear 29. Both gears are placed in a closed housing, which allows this pair to be used in parallel as an oil pump. The oil taken from the sump (not shown in Fig.) Is pumped through the channel 30 in the front end cover and the hole in the shaft 9 into the internal channel of the shaft 9. The gear 29, mounted on the shaft 31, rotates the gear 32 located in meshing with gear 33, which, in combination with pin 34, acts as a crank elbow corresponding to radius 4 (see diagram 8b). The finger 34 is included in the radial groove 35 of the gear 36, which plays the role of radius r ₁ . Gear 33 is mounted on a shaft 37 rotating in a bearing mounted in a partition (wall) 38. A wheel 39 is mounted on the second end of the shaft 37 with a pin 40 moving along the radial groove 41 in the gear 42. The wheel 39 with the pin 40 forms a second crank with the same radius r, but operating in antiphase (rotated 180 ^° ). Accordingly, the groove 41 of the gear 42 acts as a second movable radius (r ₂ ). Gear 36 is in constant engagement with gear 43 of shaft 18, and gear 42 with gear 44 of shaft 19. Shafts of gears 36 and 43 are mounted in bearings 45 and 46 mounted respectively in partitions (end walls) 47 and 48 on a common geometric axis: located offset from the axis of the shaft 37 by a.

In FIG. 8b presents several positions of variable radii r ₁ and r ₂ , rotating around the center O ₁ , offset by a value relative to the center O of rotation of constant radius r.

 In FIG. 9 is a graph of the variation in the angular velocity of the chamber shafts and the cycle duration for various numbers of pairs of blades. The graph is built in relation to compressor 2 (Fig. 6). .

In FIG. 9a graphically shows the dependences ω ₁ , ω ₂ = f (α) and α ₁ , α ₂ = f (α) for the options m = 1, m = 2 and m = 3, where ω ₁ is the angular velocity of the shaft rotating with a greater speed, ω ₂ - the angular velocity of the shaft rotating at a lower speed, α ₁ , α ₂ - the rotation angles of the respective shafts, and m - the number of pairs of blades.

 In FIG. 10 in general terms is a schematic diagram of the engine and 6 provisions of its working process (cycle).

 The engine consists of a cylindrical chamber 10 with a pair of coaxial shafts 11 installed along its axis with two pairs of opposed sector blades A-A and B-B mounted on them. The shafts of the pair 11 are interconnected by a movement conversion mechanism (not shown in the diagram). The inlet 12 and outlet 13 openings are made in the cylindrical body, as well as the hole for the spark plug 14. The cylindrical body 10, blades AA and BB, and the side walls form four chambers (I, II, III, IV), the volumes of which vary throughout the work cycle.

 In FIG. 11 are indicated: 1 - compressor; 2 - heating chamber; 3 - combustion chamber; 4 - working part; 5 - cooler working fluid.

 The principle of operation of the engine according to option I with the translational movement of the pistons and a rigid shaft (Fig. 1).

 When starting a cold engine, the accumulation or pulsating spark plug 12, which is powered by a rechargeable battery (not shown in FIG. 1), is preliminarily switched on, after which the crankshaft 4 is driven through the flywheel 19 through the flywheel 19. with the starter turned on (with separate mixture formation), the high-pressure fuel pump (not shown in Fig. 1) is turned on, which supplies fuel from the fuel tank (not shown in Fig. 1) through the nozzle 11 to the combustion chamber 10. Variation is possible m external mixture formation, ie. e. using a carburetor or other device for preparing the combustible mixture. In this case, the need for a high-pressure fuel pump and nozzle 11 disappears, however, there is a danger of self-ignition of the mixture during compression. The choice of a mixture formation system depends on the corresponding properties of the fuel on which the engine will operate.

 The rotation of the crankshaft 4 drives the piston 2 of the compressor cylinder - (compressor) 1, which, when moving downward, sucks air or combustible mixture (with external mixture formation) into the compressor 1 through the intake pipe 15 with the intake valve 16. When the piston 2 moves up the valve 16 closes, and the bypass valve 13 continues to remain in the closed position almost until the end of compression. At the end of the compression stroke, when the pressure of the compressed air (or mixture) reaches a sufficiently large value, the valve 13 opens, and the compressed air is pumped through the bypass pipe 8 into the combustion chamber 10. Since the free space above the bottom of the piston 2 is in its upper position (TDC) is zero, all or almost all of the air (mixture) is displaced from the compressor 1 into the combustion chamber 10, which removes the problem of residual gases. Upon reaching the appropriate conditions (pressure) in the combustion chamber 10, the mixture ignites from the spark plug 12, and the moment of ignition of the mixture does not matter, only the fact of ignition is important.

 When the shaft 4 rotates during engine start-up, the piston 6 of the working cylinder 5 also performs a similar movement, only in antiphase, in which the expansion and exhaust strokes occur, since the mixture is combusted in the combustion chamber 10 which is outside its limits. mixture, enters the working cylinder 5 through the bypass pipe 9 with the bypass valve 14 during the working stroke (down) of the piston 6. The exhaust working fluid from the working cylinder 5 is discharged through the exhaust pipe 17 with the exhaust to the valve 18 when the piston 6 moves up.

 In FIG. 1 shows the position corresponding to the end of compression and bypass in compressor 1 and the end of expansion and the beginning of release in the working cylinder 5.

 After igniting the mixture from the spark plug 12, the engine warming up process begins, as a result of which the combustion chamber 10 heats up, and self-ignition of the mixture injected into the combustion chamber 10 begins.

 When starting the engine until the mixture ignites, the pressure in the combustion chamber increases only due to the difference in the mass of the mixture (air) injected into the combustion chamber by the compressor 1 and discharged from it through the working cylinder 5 into the atmosphere. Due to this difference (the mass of the discharged mixture is less than the mass of the injected), an accumulation of the mass of the mixture remaining in the combustion chamber occurs. The magnitude of this difference is determined by the corresponding gas distribution phases, i.e., the time-section ratio of valves 13 and 14. The gas distribution mechanism is essentially the same as the cam profile and, accordingly, the gas distribution phases from the gas distribution mechanism of a classical internal combustion engine, therefore, in FIG. 1 is not shown. As an additional means of increasing the pressure in the combustion chamber 10, and hence the acceleration of start-up, it is possible to use a throttling of the mixture flow either in pipeline 9 or in pipe 17.

 Due to the fact that the pressure in the combustion chamber until a balance is established between the mass of the injected mixture and its mass discharged into the atmosphere increases with each supply of a new charge, the rate of increase of this pressure, ceteris paribus, depends only on the ratio of the volume of the combustion chamber 10 and the working compressor volume 1. The larger the relative volume of the combustion chamber 10, the lower the rate of pressure rise, and vice versa.

 After ignition of the mixture in the combustion chamber 10, the engine operation process begins. The fundamental differences between the operation process and the engine starting process are as follows. Firstly, the external drive mechanism is turned off, and the engine switches to the mode of operation from the thermal energy generated as a result of combustion of the air-fuel mixture. Secondly, the working pressure and temperature in the combustion chamber 10 sharply increase. Thirdly, the mixture spontaneously ignites. Fourth, the combustion process of the mixture becomes continuous with some fluctuation in the working pressure in the combustion chamber 10. Fifth, it becomes possible to switch to lean mixtures and / or other types of fuel.

Since the chemical reaction does not have enough air in the engine to completely burn the fuel, regulation of its supply (mixture composition) is, although desirable, but not necessary. For this reason, the engine is controlled only by changing the amount of fuel supply without regulating the air supply, although the option with adjustable air supply is also possible. As when starting the engine, the air (mixture) entering the compressor 1 at the inlet is compressed and pumped by the compressor into the combustion chamber 10. The pressure of the compression end, by the time the bypass valve 13 is opened, should approximately correspond to the pressure in the combustion chamber 10, although it may deviate somewhat from him. After opening the valve 13, compressed air (mixture) under the influence of its own pressure and piston 2 enters the combustion chamber 10 and immediately ignites, respectively, increasing the pressure in it by a certain amount depending on the filling of the compressor 1 and its relative (to the combustion chamber 10) volume . In FIG. 1 shows the position of the end of compression and beginning bypass inlet (TDC) in the compressor 1 and the end of expansion and start of manufacture (BDC) in the working cylinder 5. Upon subsequent rotation of the shaft 4 at ¹⁸⁰ ° piston 2 moves downward to BDC, the air sucking new serving (mixture ) At the same time, the piston 6 of the working cylinder 5 moves from BDC to TDC, pushing the spent working fluid through the opening exhaust valve 18 into the atmosphere. In the TDC position, the bypass valve 14 opens, and the working fluid under pressure in the combustion chamber 10 enters the working cylinder 5, acting on the piston 6 and moving it to the BDC. Thus, for 360 ^o rotation of the crankshaft 4, a complete cycle is performed with the separation of the intake-compression and expansion-exhaust cycles between the compressor 1 and the working cylinder 5, as well as taking the combustion process beyond. The increase in pressure in the combustion chamber 10 when applying the next charge of the mixture is leveled off by a corresponding drop in it when the next portion of the working fluid is fed into the working cylinder 5, therefore, almost at a steady state, the working process proceeds with constant pressure with insignificant pulsation.

 The main advantages of the proposed engine from the general analysis of the PV diagram shown in FIG. 2. To ensure the most correct approach to the comparative assessment of both options, a number of preconditions were adopted. First of all, the condition of equality of the working volumes of the cylinders of the compared variants was adopted, and in the case of the proposed engine, the working volume of the cylinder is also its total volume.

Since the volume of the compression chamber, as part of the cylinder volume of the proposed engine, is practically zero, and the standard internal combustion engine with the adopted compression ratio Σ = 9 is approximately 11% of the total cylinder volume, the latter has a shorter piston stroke. To comply with the condition of equality of working volumes in the calculations, the cylinder diameter of the reference internal combustion engine is adopted increased by 6%. Finally, if the cycle of the proposed engine takes place in two cylinders (compressor and slave cylinder), then in the case of a reference internal combustion engine - only in one cylinder. Thus, if the working volumes of the compressor, the working cylinder and the cylinder of the standard internal combustion engine are equal, the total working (also full) volume of the proposed engine will be twice the working volume of the standard internal combustion engine. But in the first case, the full cycle is completed for 360 ^{about the} rotation of the shaft, and in the second for 720 ^about . Hence, one cycle of the proposed engine accounts for half of the cycle of the internal combustion engine, which ensures compliance with the principle of equality of conditions (volumes).

 The PV diagram shows all the cycles of a four-stroke engine, with the only difference being that in the case of a reference internal combustion engine they belong to one cylinder, and in the case of the proposed engine, they are distributed in pairs between the compressor and the working cylinder.

When calculating and constructing the PV diagram, the following initial parameters and their values were taken: initial compression pressure P _a = P _ao = 0.8 kgf / cm ² ; compression ratio Σ = 30, Σ _o = 9; the degree of expansion -Σ = 6.7 and 5 (second option); - Σ _o = 9; the compression polytropic index n ₁ = 1.35; combustion pressure P _Zo = 60 kgf / cm ² , P _Z = 86 kgf / cm ² .

The equality of the values P _a, r. F.. P _a = P _{ao is} accepted only to comply with the principle of equality of initial conditions for both compared options. In fact, P _a ≠ P _ao due to the difference in temperature conditions, the presence of residual gases in the reference internal combustion engine, etc.

= ∞ .The compression ratio Σ for the proposed engine is a value representing the ratio of the total volume of the compressor to its volume at the point P _c , i.e., at the end of compression and the beginning of the bypass. The degree of expansion is a value equal to the ratio of the total volume of the working cylinder to its volume at the point P _Z ^I corresponding to the end of the bypass of the working fluid from the combustion chamber to the working cylinder and the beginning of the expansion process in its pure form. For the standard internal combustion engine, the degree of expansion is adopted - Σ _o = 9. For the proposed engine - Σ = 6.7 and - Σ = 5 (second option). In fact, expansion begins at point P, with zero volume V = 0, i.e., - Σ =

= ∞.

For the compression polytropic exponent n ₁ in both cases, the average value n ₁ = 1.35, typical of modern ICEs, is accepted.

 The coordinates of the characteristic points of the PV diagram were obtained by calculation on the basis of the given initial data taken from the technical literature.

 The given PV diagram is mainly illustrative and does not claim absolute accuracy.

 In the analysis of the entire working cycle, taking into account the above comments, an ideal cycle is considered mainly.

Intake process. For both compared options, this process begins at TDC and TDC _about , i.e., at the end of the release process. The difference is that in the case of the proposed engine, the intake is made into the compressor, and in the case of the reference internal combustion engine - into the same cylinder in which the entire cycle is completed. The presence of a combustion chamber in the cylinder of the standard internal combustion engine determines the presence of residual gases in it, which results in a decrease in its filling with fresh air-fuel mixture (or air) and a deterioration in the composition of the mixture (air) due to the admixture of non-combustible residual gases (up to 11%). Although taken equality condition that the initial intake pressure P _a made the same for both cases, ie. F.. P _a = P _ao , in fact, the presence of residual gases significantly increases it, and if, in the case of the proposed engine, the inlet starts at zero or close to zero pressure, then in the case of a reference internal combustion engine, the pressure at which the inlet starts is close to atmospheric or exceeds it. An additional factor that reduces the filling of the cylinder is its higher temperature in the case of a reference internal combustion engine. This general process is completed for both NMT and embodiments described in the diagram the curves TDC-BDC and TDC _of -NMT, respectively, at equal pressures P _a = = P _ao = 0,8 kgf / cm ^2.

Compression process. Compression starts at BDC common to both embodiments at a pressure P _a = P _ao = 0,8 kgf / cm ^2. It is similar for both options, but in the details it has very significant differences. In the proposed engine, it occurs in the compressor and continues in its pure form to the point P _c , the abscissa of which corresponds to about 2.5-3.5 of the volume of the latter, and the ordinate - about 80 kgf / cm ² . The completion of the compression process is the bypass described by the section P _c -P _{Z of the} curve P _a -P _c -P _Z with the abscissa of the point P _Z equal to zero and the ordinate equal to approximately 86 kgf / cm ² . In this case, the bypass is similar in nature to compression, but unlike the latter, it is accompanied by a counterflow of hot (or burning) working fluid from the combustion chamber to the compressor, but only if the pressure at point P _{c is} lower than pressure P _Z. In this case, the counterflow is instantaneous in nature, since pressure equalization immediately sets in, that is, the condition P _c = P _Z , after which the counterflow ceases. If the pressure P _c ≥ P _Z , no counterflow occurs.

In contrast to the proposed engine in the standard internal combustion engine, the compression process takes place in the same cylinder as the entire cycle, and ends in its pure form at the point P _{co the} general compression curve P _ao -P _co -P _Zo with an abscissa corresponding to about 0.2 full and 0.1 of the working volume of this cylinder. The completion of the overall compression process is a mixed process-combustion-compression process, the described section P _co -P _{Zo of the} overall compression curve. These differences affect the characteristics of the process and, as a consequence, the output parameters of the engine. In the standard internal combustion engine, it proceeds at a higher temperature in general, and on P _co -P _Zo - with a sharp increase in temperature and pressure due to ignition and combustion of the mixture charge, i.e., with high back pressure, and therefore a large reactive moment, to overcome which a significant part of the power is wasted. The power of the proposed engine is spent only on compressing the charge of the mixture (air), because even in a small area P _c -P _{Z of the} general compression curve, i.e., when bypassing, this additional power consumption is used to increase the pressure in the combustion chamber. The lower process temperature can significantly increase the degree of compression when using a fuel-air mixture (external mixture formation). The completion of the compression process (including bypass) at the point P _Z with an abscissa corresponds to the zero volume of the cylinder (V = 0) and with the ordinate corresponding to the pressure in the combustion chamber completely eliminates the problem of residual gases, provides a smooth transition from the pressure of the end of compression to the pressure of the working fluid in the combustion chamber and the maximum use of the charge of the mixture (air).

 The process of combustion of the mixture. The removal of the combustion process of the mixture beyond the working space of the proposed engine provides him with several advantages. In addition to the advantages mentioned, it should be added that due to the supply of a fresh charge to the atmosphere of high-temperature gases already burnt or continuing to burn, the problem of ignition of the fresh charge mixture and its complete combustion does not arise. In contrast to the standard internal combustion engine, this applies to mixtures of a wide range both in composition (coefficient of excess air α) and in types and grades of fuel.

Expansion process. The main differences in the characteristics of this process for the compared options are due to the fact that in the case of the proposed engine, it occurs in a separate working cylinder, and in the case of a reference internal combustion engine - in the cylinder common to the entire cycle. The advantage of the proposed engine is that in the first case, the expansion begins with a bypass at zero volume of the cylinder (point P _Z with an abscissa corresponding to TDC) and continues at constant pressure to point P _Z ^I (section P _Z -P _Z ^{I of the} general curve extensions). The constancy of pressure in this section is ensured by the balance of the supply of the mixture or air and fuel (with separate mixture formation) into the combustion chamber 10 (Fig. 1) and the flow rate of the working fluid during bypass. The absolute value P _Z = P _Z ^I for each specific mode, the load is regulated by the flow similar to the diesel process. In contrast to the reference internal combustion engine, here at the initial stage of the process, the cylinder volume determined by the abscissa of the point P _Z ^{I is} additionally used. In this example, it makes up about 15% of the cylinder volume.

In the case of a reference internal combustion engine, expansion begins at point P _Zo with an abscissa corresponding to the volume of the combustion chamber, i.e., approximately 11% of the total volume of the cylinder. Consequently, up to 11% of the cylinder volume is not involved in the expansion process.

In the case of the proposed engine, an already prepared working fluid is involved in the expansion process, which is the product of complete or close combustion of the mixture, therefore, after bypassing (at the point P _Z ^I ), expansion in its pure form begins. This allows, by appropriate selection of the bypass value (segment length P _Z -P _Z ^I ), to establish (during design) the expansion depth, i.e., the degree of utilization of the energy of the working fluid, and thus determine the power-efficiency ratio. With an increase in the length of the segment P _Z -P _Z ^I , the power increases, with a decrease in efficiency.

Theoretically, it is possible in this way to bring the expansion process to a cold outlet, i.e., to maximize the use of the energy of the working fluid by sacrificing output power, or to bring the power to a maximum limited only by the amount of backpressure of the outlet. In FIG. 2, an option is additionally presented with an increased length of the segment P _Z -P _Z ^I to P _Z -P _Z ^II , giving a corresponding increase in the area of the figure, describing the operation and, as a consequence, an increase in engine power.

In the case of a reference internal combustion engine, such a dosing of the flow of the working fluid is excluded, in addition, the expansion process in it is practically represented by three complex processes — ignition — preliminary combustion of the mixture (section P _co -P _Zo ), combustion of the mixture (section P _Zo -P _Zo ^I ) and extension-completion. In this case, the formation of the working fluid occurs at all of the stages listed and does not completely end or ends outside the engine proper in the afterburner, that is, it is useless. The process of expansion of both ends in the compared options common to them at the ordinate P _a = P _ao and the abscissa corresponding to BDC.

Release process. Described in PV diagram, the curves having a common origin with the abscissa and ordinate BDC P _a = = P _ao and ends at _about TDC and TDC, respectively. However, this is true only for an ideal cycle. In a real cycle, production begins earlier than TDC (TDC _o ) to reduce backpressure of the combustion products and, accordingly, to more completely clean the cylinder of combustion products. The beginning of the release with reference to the internal combustion engine is the beginning of the opening of the exhaust valve (point P _o ), i.e., at a pressure of about 4-8 kgf / cm ² , which entails a significant loss of energy and noise, for which the silencers of the exhaust noise are used to suppress, creating additional resistance and additional energy loss. Moreover, in modern internal combustion engines, devices such as afterburners, catalytic and thermal reactors are used to reduce the toxicity of combustion products, which, creating even greater resistance to exhaust, require the enrichment of combustion products and the supply of secondary air to the exhaust duct for normal operation. All this leads to large energy losses and reduced fuel efficiency, as well as power loss due to the significant value of the coefficient of residual gases.

In the case of the proposed engine, it is possible to avoid these disadvantages: due to the completeness of the combustion of the mixture, environmentally friendly combustion products enter the exhaust; the possibility of dispensing the supply of the working fluid to the working cylinder allows P _a to be minimized (theoretically below atmospheric pressure), i.e., to eliminate the need for exhaust silencing devices; clean release eliminates the need for devices to neutralize the toxic components in it. All this simultaneously increases the environmental friendliness and fuel efficiency of the engine and brings them to an acceptable optimum.

 A significant part of these advantages is clearly visible when comparing the areas of the figures corresponding to the work of the compared options in one cycle of the same working cylinder volumes, if we take the volume of the working cylinder of the proposed engine for its working volume.

 The principle of the engine according to option II - improved option 1 (Fig. 3). The main difference between option II and option I is the separation of the common crankshaft 4 (Fig. 1) into two crankshafts - 21 (compressor shaft 1) and 28 (working cylinder shaft 5) and output shaft 23 in combination with the introduction of a planetary mechanism (differential ) 19 (Fig. 3). The introduction of these improvements creates a fundamentally new effect.

 Starting a cold engine, in principle, is the same as in option 1, however, it has a number of differences due to its circuit.

, и составляющим в данном случае (при d_к = d_в) i_о = 1. Это означает, что при некоторых условиях (например, пуск двигателя стоящего автомобиля с включенной трансмиссией), когда крутящий момент на выходном валу 23 недостаточен для проворачивания последнего, вал 23 можно считать заблокированным (неподвижным). В этом случае вал 21 компрессора 1 будет вращаться с удвоенной частотой по отношению к частоте вращения водила 24 и вала 28 рабочего цилиндра 5, если принять передаточное число для зацепления пары шестерня 27 - зубчатый венец 26 равным 1. Распределение момента соответствует i_о. При этом подача смеси компрессором 1 в камеру сгорания 10 будет осуществляться с удвоенной частотой, что значительно сокращает время пуска и облегчает сам пуск. После воспламенения смеси в камере сгорания 10 резко возрастает давление, что приводит к увеличению реактивного крутящего момента (-М) на валу 21 компрессора 1 и, соответственно, активного момента (М_р) на водиле 24. Одновременно возрастает и момент М (М = -М) на выходном валу 23.Starting the engine from an external drive (starter) is carried out by rotating the carrier (differential box) 24. In this case, the torque on the carrier 24 is distributed between the gears 20 and 22 of the shaft 21 of the compressor 1 and the output shaft 23 in accordance with the internal gear ratio i _о , which is the ratio of the diameter d _{to the} gear of the shaft 21 of the compressor 1 to the diameter d _{b of the} gear 22 of the output shaft 23 (Fig. 3), i.e., i ₀ =

, and components in this case (with d _к = d _в ) i _о = 1. This means that under certain conditions (for example, starting the engine of a standing car with the transmission engaged), when the torque on the output shaft 23 is insufficient to rotate the latter, the shaft 23 can be considered locked (fixed). In this case, the shaft 21 of the compressor 1 will rotate at twice the frequency with respect to the rotational speed of the carrier 24 and the shaft 28 of the working cylinder 5, if we take the gear ratio for engagement of the pair of gear 27 - ring gear 26 equal to 1. The distribution of torque corresponds to i _about . In this case, the mixture is supplied by the compressor 1 to the combustion chamber 10 with a double frequency, which significantly reduces the start time and facilitates the start itself. After ignition of the mixture in the combustion chamber 10, the pressure increases sharply, which leads to an increase in the reactive torque (-M) on the shaft 21 of the compressor 1 and, accordingly, the active moment (M _p ) on the carrier 24. At the same time, the moment M (M = - M) on the output shaft 23.

After starting the engine, when the heating is turned on, the combustion of the mixture in the combustion chamber 10, as in the case of option 1, is stabilized, i.e., it becomes continuous. In this case, the distribution of moments -M and M, as components on the carrier 24, remains the same as at start-up. However, if during start-up the moment on carrier 24 is created by an external device (starter), then when the engine is running, by the working cylinder 5 as a result of using the energy of the combustion products of the mixture. This changes the nature of not only the origin, but also the distribution of torque. First of all, if at start-up the moment supplied from the outside is spent only on overcoming the resistance to rotation of the shafts 21 and 23, then when the engine is running, the corresponding part of the moment created by the working cylinder 5, i.e., the moment M _p , is also spent on creating the latter (compressor drive 1, providing the mixture in the combustion chamber). By analogy with option 1, there are significant differences. The first of these consists in increasing the speed n _{to the} shaft 21 of the compressor 1 during start-up. The second - in self-regulation of the ratio of rotational speeds n _p drove and n _to - the compressor shaft. The third - in the possibility of the engine working "on itself", that is, when rotating only the shaft 21 of the compressor 1 and the shaft 28 of the working cylinder 5 with a stationary output shaft 23.

Upon reaching the output shaft the moment M is sufficient to drive it into rotation, its rotation frequency n becomes non-zero, and as it increases, the quantity n _k decreases, and therefore the moment M _p developed by the working cylinder 5. When the value decreases n _in relatively increases the value of n _to , which leads to an increase in the absolute value of M _p , and therefore to an increase in the absolute value of the output moment M in accordance with the value of i _about . Ultimately, the magnitude of the output torque M of the engine within the range of its changes, determined by the value of i _o , becomes a function of n. From here, we can write: M = f (kn ^-1 i _о ), where k is the coefficient of correspondence of moment M to frequency n, depending, in particular, on a number of factors determined experimentally.

 Since the process of changing the value of M depending on n is continuous, this leads to its self-regulation in the appropriate range depending on specific conditions (for example, resistance to movement, in the case of a car), and to optimization of the engine operating mode. This means that, unlike the standard internal combustion engine, this engine automatically reaches the optimal mode (maximum torque mode for any load), i.e., the maximum energy recovery mode for the working fluid. Moreover, at any feed rate, i.e., independently of it, the engine develops the moment M necessary to overcome the resistance to movement (in the case of a car), up to the maximum possible value at the corresponding speed n of the output shaft. It can develop the maximum possible moment even in idle mode at the corresponding (minimum) frequency n.

You can set any value i _o , n. If the preference is given to power, i _o > 1 is accepted, if profitability, i _o <1 is accepted.

In FIG. 4 a family of curves presents the dependence M = f (i _o ). The abscissa axis represents the values of n, and the ordinate axis represents the values of M as a percentage of its value, taken as 100% at i _o = 0.5.

All five curves presented have a similar character and typical inflection points n ₁ , n ₂ , n ₃ , n ₄ , n ₅ , corresponding to the beginning of a sharp decrease in the value of M and n, due to the exhaustion of the reserve M to overcome internal resistance in the engine.

The function M = f (i _o ), presented on the graph, shows a direct dependence of M and, accordingly, power on i _o . So, a 2-fold increase in i _о (from i _o = 0.5 to i _о = 1) gives a corresponding (up to 200%) increase in M. With an increase in M due to an increase in i _о , the total fuel consumption also increases.

When analyzing the graph, at first glance, an idea arises about its mismatch of the inverse dependence of M on n. So, the curve i _o = 1.2, corresponding to the maximum (for this example) value of M = 240%, has an abscissa of the end point of the order of n = 11600 min ^-1 . However, this dependence is valid only for each specific quantity i _о . Within the same curve, an increase in n from 0 to 11,600 min ^-1 gives a decrease in M from 340 to 0%. The curve i _o = = 0.3 gives a decrease in M from 60 to 0%.

As for the direct dependence of n on M presented on the graph, it is explained by the fact that a larger initial moment (at a zero value of n) provides a larger final moment, which leads to the extension of the corresponding curve. However, within each individual curve, this inverse relationship is preserved. The value of i _about determines the torque conversion coefficient.

Technically, the value of i _о ≠ 1 is achieved by the use of a device such as planetary mechanisms, which allow to divide the rotation frequency at the input into two different frequencies at the output. In this case, it is proposed to use the differential for this purpose, as a special case of the planetary mechanism.

As can be seen in the circuit diagram of this device shown in FIG. 5, it is a differential with bevel gears 20 of the crankshaft 21 of the compressor and 22 of the output shaft 23. The gears are different in diameter, which gives i _о ≠ 1. In this case, d ₂₀ > d ₂₃ , therefore i _о > 1. Replacing gear 20 and 22 places allows you to get the inverse of i _o , that is, i _o <1. The gears mesh with the satellites 25 'and 25, respectively. When the carrier 24 rotates around its own axis, the satellites 25 ^I and 25, rigidly mounted on one common shaft installed in the bearings of the carrier 24, will distribute the moment on carrier 24 between the gears 20 and 22 in direct proportion to their diameters.

 Option III - RPD with a planetary mechanism (Fig. 6) is a rotary piston equivalent of options II and 1, as a simpler technical solution that differs from option II by the absence of a planetary mechanism. As shown in FIG. 6, option III includes all the basic elements of option II, as well as several additional mechanisms and devices.

The rotation during start-up by the starter (not shown in Fig. 6) drove the planetary gear 26 (or differential) 8 through the satellites to the sun gear 17 and the shaft 9 rigidly connected to it, which transmits it to the input of the control mechanism 7 of the compressor 2 located in the compartment 1, where it is converted from a uniform rotation of the shaft 9 to the rotation of the coaxial shafts 18 and 19 with a sinusoidal change in angular velocities in antiphase. This leads to a change in volumes enclosed between the blades 21 and 20, rigidly mounted on the shafts 18 and 19, respectively. Since the blades 20 and 21 divide the internal space of the compressor 2 into two parts, forming two mutually changing angles, a decrease in one of them entails a corresponding increase in the other, and vice versa. In compressor 2, 2 processes occur simultaneously. In that part of the chamber where the angle, and therefore the volume between the blades 18 and 19 increases, air (or air-fuel mixture) is sucked (inlet) through the inlet pipe 13 with a throttle or air damper. In the part of compressor 2 with a decreasing volume, compression occurs, starting from the moment of blocking the inlet window by the body of one of the blades (20 or 21), rotating at a lower speed, and ending at the beginning of the bypass, i.e., at the point P _c - on the PV diagram ( Fig. 2). During the bypass, air (or the air-fuel mixture) is pumped into the combustion chamber 3 through the bypass window in the end wall between the compressor 2 and the combustion chamber 3. In the combustion chamber 3, the air is mixed with the fuel supplied through the nozzle 15 (in the case of separate mixing), and obtained ( or finished) the mixture is ignited by candle 16.

In parallel, the rotation of the carrier 26 through the hollow shaft 11 which is rigidly connected with it and the gear system is transmitted to the input of the regulating mechanism 5 of the working chamber 4, where from the uniform rotation of the shaft 11 it is converted at the output into the rotation of the coaxial hollow shafts 24 and 25 with the blades 22 and 23 fixed to them with a sinusoidal change in angular velocities in antiphase. As in the case of compressor 2, this leads to a change in the angles, and therefore the volumes of the working chamber 4, enclosed between the blades 22 and 23, which creates the necessary conditions for the expansion and exhaust processes that occur simultaneously with the intake and compression processes in the compressor 2. The presence of the planetary mechanism 8 creates the necessary conditions for the rotation of the shafts 9 and 11 with different frequencies n _k and n _p in accordance with the distribution of moments, however, this is manifested only in the duration of the processes and does not exclude the simultaneity of their flow Niya. The supply of air (mixture) by the compressor 2 to the combustion chamber 3 is accompanied by the simultaneous exit of this air (mixture) from it into the working chamber 4 through the bypass window in the wall between the combustion chamber 3 and the working chamber 4, and then, through the exhaust stroke in the working chamber 4, through exhaust pipe 14 to the atmosphere. As in the case of options I and II, the pressure of the mixture in the combustion chamber 3 increases due to the difference in its supply and exhaust.

 After igniting the mixture in the combustion chamber 3 from the candle 16 and warming up the combustion chamber 3, the process of continuous combustion starts with self-ignition of the mixture, and the starter and power supply to the spark plug 16 are turned off.

 When starting the base, the function of the combustion chamber 3 is reduced to the accumulation of the mixture or its formation and accumulation, i.e. it is of a preparatory nature, then after ignition of the mixture and the start of engine operation, the combustion chamber 3 additionally acquires the functions of burning the mixture and forming a working fluid, i.e. functions of the furnace and boiler. Its combination with the compressor 2, acting as a kind of injector, forms a kind of generator of the working fluid that creates the working pressure.

 Before the engine starts, the working chamber 4 acts as a pumping pump that pumps the mixture from the combustion chamber to the atmosphere. After ignition of the mixture and the beginning of its steady burning, the working chamber 4 turns into the actual engine, working under the pressure of the working fluid and creating the appropriate torque.

 Starting the engine was considered with reference to the stationary output shaft 12, connected with the third element of the planetary mechanism 8, i.e., the crown. After starting and starting the engine, the shaft 12 remains stationary until the torque M reaches the value necessary to drive it into rotation, after which the relative speed n of the shaft 12 (relative to the shaft 9 of the compressor 2) begins to increase, and the frequency of the shaft 9 decreases, which leads to a corresponding decrease in the value of M. It should be noted that we are talking only about a relative, and not an absolute decrease and increase in the corresponding rotational speeds and moment M. Absolutely they can increase.

 In FIG. 7 presents the technical solution of the compressor chamber 2 and circuit diagrams of the compressor 2 and the working 4 engine chambers (Fig. 6). Since both chambers have a generally similar design and differ mainly only in the direction of flows and the nature of the gases circulating in them, it suffices to confine ourselves to a description of the technical solution of one of them, for example, compressor (compressor) 2, and show the differences on the schematic diagrams. The technical solution of the camera is presented in the simplest, two-bladed (with one pair of blades) version (Fig. 7a).

The rotation of the shaft 9 (see also Fig. 6), converted by the regulating mechanism (Fig. 8) into the rotation of the shafts 18 and 19, and therefore the blades 21 and 20, respectively, with mutually (in antiphase) changing angular velocities leads to a change in angles , and therefore two parts of the volume of the chamber 2, enclosed between the blades. The sum of the angles that the shafts 18 and 19 rotate (α ₁₈ and α ₁₉ ), and therefore the blades 21 and 20, is α ₁₈ + α ₁₉ = α ₉ = const, where α ₉ is the angle of rotation of the shaft 9. Since in this case the range of changes of α ₉ is in the range 0 ^about <α ₉ <360 ^about , and α ₁₉ -0 ^about <α ₁₉ <(360 ^about -α ₁₈ ). Over the next cycle, the ranges of changes of α ₁₈ and α ₁₉ are mutually changed, and the blades 20 and 21 are mutually interchanged. The blades 20 and 21 are in the form of sectors, that is, they occupy some part of the internal volume of chamber 2 with their bodies, therefore α ₁₈ and α ₁₉ represent the angles of rotation of the bisectors of these sectors. Theoretically, the blades 20 and 21 can have any thickness (the angle of the sectors forming them), but under the conditions of equality of these angles, and also that their total angle should not exceed 180 ^about . Reducing the angles of the sectors of the blades, respectively, increases the free space of the camera, i.e., its working volume. However, two limiting factors arise here. The first of them is due to strength limitations, since a decrease in the angle of the sector of the blade entails a decrease in thickness, and hence its strength at the base. The second is associated with the gas distribution phases, since the body of the blades is used to block the inlet 35 and the bypass 36 windows, and therefore its dimensions should be consistent with the size and shape of the windows. In the case under consideration, the maximum possible thickness of the blades was adopted, i.e. 90 ^° , which in the case is 180 ^° . Hence, for the cycle, the rotation angles of the shafts are α ₉ = 360 ^о , α ₁₉ = 270 ^о , α ₁₈ = 90 ^о . At each subsequent cycle, the values of α ₁₉ and α ₁₈ mutually change.

The diagram (Fig. 7) shows 3 positions of the cycle corresponding to its various stages (values α ₉ ). Position I corresponds to TDC, i.e., α ₉ = α ₁₈ = α ₁₉ = 0 ^about . In this case, the instantaneous angular velocities ω of the shafts 18 and 19 are equal (ω ₁₈ = ω ₁₉ ), the blades 20 and 21 are brought to the limit, the position of the blade 20 corresponds to the moment the inlet window 35 begins to open, and the blade 18 completely closes the bypass window 36.

When the shaft 9 is rotated clockwise by a certain angle, the shafts 18 and 19 rotate in the same direction, however, the shaft 19 is accelerated and the shaft 18 is decelerated. This entails an increase in the gap (angle) between the blades and, as a result, an increase in the volume of the inner space of the chamber enclosed between them. At the same time, the blade 20 begins to open the inlet window 35, and the blade 18 closes it, but more slowly, and also continues to close the bypass window 36. As a result of the resulting vacuum in the space between the blades 18 and 19, the air (mixture) is admitted through the window 35. In position II (α ₉ = 180 ^about , α ₁₉ = 135 ^about , α ₁₈ = 45 ^about ) corresponding to half the cycle, the inlet window 35 is fully open by the blade 20, but half closed by the blade 21, which continues to close the bypass window 36. This the position also corresponds to a maximum of ω ₁₉ and a minimum of ω ₁₈ . With a further increase in α _{9, the} difference between the angular velocities ω ₁₉ and ω ₁₈ increases, which leads to an even larger increase in the angle, and hence the volume of space between the blades, i.e., to an increase in the flow of air (mixture) into the chamber. In this case, the bypass window 36 continues to remain closed by the blade 21. In position III (α ₉ = 351 ^about , α ₁₉ = 265 ^about , α ₁₈ = 86 ^about ) the inlet process in its pure form ends. Complete rotation of the shaft 9 to complete rotation (α = = ³⁶⁰ degrees) accompanied by an increase α ₁₉ and α ₁₈ to the maximum, ie. E. To α ₁₉ = ²⁷⁰ ° and α ₁₈ = ^90, which corresponds to the maximum gap between the blades 21 and 20, equivalent to the working (also full) camera volume. Moreover, the whole system comes to its original position with the only difference that the places and behavior of the shafts 18 and 19 of the blades 21 and 20 are mutually changed and the next cycle begins.

At the same time, in the second part, the volume of which during the cycle adequately decreases from the maximum at α ₉ = 0 ^о to zero at α ₉ = 360 ^о , a compression process occurs, which ends with a bypass. The beginning of the bypass corresponds to the beginning of the opening of the inlet window by the blade 20. For almost the entire length, compression occurs when the inlet 35 and bypass 36 windows are closed and only at the end, a few degrees before the TDC (position III), the blade 21 with its trailing edge begins to open the bypass port 36, which corresponds to the beginning of the bypass ending at the TDC (position 1). At the end of the cycle (α ₉ = 360 ^° ), the planes of the surfaces of the blades completely (or almost completely) close, which leads to the complete displacement of air (mixture) between the blades into the combustion chamber and, as a result, gives maximum use of the compressor working volume and removes the problem of residual gases.

 Relying on the technical solution of compressor 2, it is possible to describe the working chamber 4 (Fig. 6) by restricting itself to considering its circuit diagram as applied, as in the case of compressor 2, to those basic provisions shown in Figs. 7 c. The working chamber 4 is designed for expansion and release cycles occurring simultaneously.

Position 1 (TDC) - the moment of the start of the bypass in the bursting part and the release in the tapering part of the chamber 4 (Fig. 6). The blade 22 begins to open the bypass window 37 in the partition between the combustion chamber 3 and the working chamber 4, and the blade 23 - the exhaust window 38 in the cylindrical wall of the chamber 4. The angular velocities ω of the shafts 24 and 25, and therefore the blades 22 and 23 are equal, t i.e., ω ₂₄ = ω ₂₅ , however, from this moment on, the shaft 24 with the blade 22 starts to rotate with acceleration, and the shaft 25 with the blade 23 starts to slow down, which leads to the appearance and increase of the gap (angle) between the surfaces of these blades.

 When the system rotates, the working fluid through the opening window 37 enters under pressure from the combustion chamber 3 into the increasing space between the blades 22 and 23 and, acting on them, rotates the shafts 24 and 25, which is converted by the regulating mechanism 5 into the rotation of the common shaft 9, distributed planetary mechanism 8 between the compressor 2 and the output shaft 12. The pressure of the working fluid acts on the blades in opposite directions, as if pushing them apart. However, due to the corresponding distribution of the moment between the shafts by the regulating mechanism 5, the active moment on the shaft 24 exceeds the reactive moment on the shaft 25, which leads to the rotation of both shafts in one direction. At the same time, through the opening blade 23 of the part of the outlet window 38, from the tapering part of the chamber 4, the working fluid spent for the previous cycle is released.

Position II corresponds to the moment the bypass window 37 is completely blocked by the blade 23, that is, the bypass is completed and the expansion begins in its pure form when the exhausted working fluid continues to be released through the open part of the outlet window 38. The bypass ends at a gap angle between the adjacent surfaces of the blades 22 and 23, equal to 22.5 ^about , which corresponds to the abscissa of the point P _Z ^I on the PV diagram (Fig. 2) or 12.5% of the volume of the chamber 4 (Fig. 6).

 Changing (when designing) the size of the bypass window 37 (the angle of the window sector) allows you to change the dose of the working fluid entering the chamber 4, therefore, increase or decrease the work performed by the engine and the power developed by it. The increase in work (increase in power) leads to increased consumption of the working fluid, which requires an increase in the supply of the mixture (or fuel and air) to the combustion chamber 3, which is achieved, in particular, by increasing the number of cycles of the compressor 2 by one cycle of the working chamber 4. When further rotation of the shaft 9, and therefore the shafts 24 and 25, the process of pure expansion, accompanied by the release continues.

350^о, α₂₂

262^о, α₂₃

87^о) наступает завершающая стадия расширения и выпуска, после чего система приходит в положение (α₉ = 360^o, α₂₂ = 270^о, α₂₃ = 90^о), соответствующее концу цикла и началу следующего цикла. Оно аналогично положению 1, за исключением взаимозамененных мест лопастей 22 и 23.In position III (α ₉

350 ^about , α ₂₂

262 ^about , α ₂₃

87 ^o ) the final stage of expansion and release begins, after which the system comes to the position (α ₉ = 360 ^o , α ₂₂ = 270 ^o , α ₂₃ = 90 ^o ), which corresponds to the end of the cycle and the beginning of the next cycle. It is similar to position 1, except for the interchangeable places of the blades 22 and 23.

 Since the principle of operation of the compressor 2 and the working chamber 4 are similar, it is enough to dwell on a more detailed consideration of the technical solution of the chamber, based on the partially considered compressor variant (Fig. 7a). Designations of details of a variant of the working chamber are also given in accordance with FIG. 6, but in parentheses. The technical solution of the camera is clearly visible in FIG. 7a. This technical solution is given in relation to the simplest, two-bladed version (with one pair of blades).

или в упорядоченном виде
ω₁=

.The principle of the mechanism is based on a change in the size (length) of the moving radius r ₁ (variable length), conjugated with a radius r (constant length) when both radii rotate around their O ₁ and O centers, spaced by a value (Fig. 8b). The ratio of the lengths of the radii r and r ₁ gives an inverse relationship of their angular velocities. Taking r = const, we can obtain the angular velocity ω ₁ (radius r ₁ ) relative to the angular velocity of radius r, i.e., the value 1 / r ₁ . In expanded form, this dependence can be represented as
ω ₁ =

or in an ordered manner
ω ₁ =

.

Thus, if you rotate the radius r around the center O, then this will entail the rotation of the radius r ₁ associated with it around its center O ₁ . Moreover, each position r in the rotation angle α corresponds to a certain position r ₁ with an angle α ₁ . As seen in the figure, with increase α in the range 0 ^{to 180} ≅α≅ radius r ₁ increases from _1min r = r - a to r _1max = r + a. In the range 180 ^о ≅α≅ 360 ^{о, the} value of r ₁ decreases in the inverse relationship. This leads to a corresponding decrease in its angular velocity ω ₁ in the range of 0 ^about -180 ^about and its increase in the range of 180 ^about -360 ^about α.

In this case, a description is given of the interaction of two radii, r and r ₁ . However, if we extend the radius r to a diameter and relate it in a similar way to a radius r ₂ rotating around the same center О ₁ , the angular velocity ω _{2 of} radius r ₂ will change in antiphase with respect to ω ₁ . This principle underlies the technical solution presented in FIG. 8a.

 When the shaft 9, and therefore the gear 28 and the gear 29 which is engaged with it, rotates, the shaft 31 is rotated with the gear 32, which is meshed with the gear 33 and sitting on the same shaft 37 with the wheel 39. The gear 33 with the finger 34 and a wheel 39 with a finger 40, playing the role of cranks operating in antiphase and defining gears 36 and 42 with radial grooves coupled with them, rotating with angular velocities mutually changing in antiphase along a sinusoid. This rotation is transmitted by the shaft 18 and 19 through the gears 43 and 44 of these shafts, which are meshed with gears 36 and 42, respectively.

Thus, the considered regulatory mechanism allows you to convert the rotation of the shaft 9 of the engine with a constant angular velocity ω into the rotation of the camera shafts with angular velocities ω ₁ and ω ₂ changing in antiphase. The presence in the engine of two mechanisms 5 and 7, i.e., for the working chamber 4 and compressor 2, in combination with the planetary mechanism 8 (Fig. 6) also allows to obtain different rotational speeds of the shafts of the compressor 2 and the working chamber 4, i.e. the possibility of self-regulation of torque and, accordingly, engine power.

The nature of these changes is shown graphically in FIG. 9. As shown in FIG. 9, in the simplest version (m = 1), i.e., with one pair of blades, for one revolution of the shaft 9, (α = 360 ^o ), one cycle of the corresponding duration (360 ^o ) is performed. Moreover, α = α ₁ + α ₂ = const, ω = const, and ω ₁ and ω ₂ vary in magnitude from ω ₁ = ω ₂ for α = α ₁ = α ₂ = 0 to ω ₁ = max, ω ₂ = min at α = = ^180, α ₁ = ^135, α = ₂ and again ^about 45 to ω ₁ = ω ₂ when α = 360 ^o, α ₁ = 270 ^° and α ₂ = 90. Thus, for ^a 360 one cycle takes place. In the case of m = 2, ie during two vanes pairs, the conditions for the two cycles per 360 ^° or one cycle for every 180 ^o (α = = 180 ^o, 360 ^°;.. Α ₁ = ^{135, 180;} α ₂ = 45 ^about , 180 ^about ).

Similarly, when m = 3 per one revolution ⁽³⁶⁰ degrees) the shaft 9, three cycles 120, and so ^on. D.

, где m - количество пар лопаток. Однако, помимо количества (пар) лопаток необходимо обеспечить выбор соответствующих значений α₁, α₂ = f(α). Технически же это достигается выбором соответствующих передаточных отношений пар шестерен 36-43 и 42-44 (фиг. 8).In general terms, the duration of one cycle in the rotation angles α of the shaft 9 can be represented as the dependence α =

where m is the number of pairs of blades. However, in addition to the number (pairs) of blades, it is necessary to ensure the selection of the corresponding values of α ₁ , α ₂ = f (α). Technically, this is achieved by selecting the appropriate gear ratios of pairs of gears 36-43 and 42-44 (Fig. 8).

 The operation of the regulating mechanism is shown with reference to the compressor, but in principle it is applicable to the working chamber.

 An example of a partial analogue is the rotary vane internal combustion engine shown in FIG. 10. The principle of operation of this engine is considered in relation to the six positions occupied by it during the working cycle. In the initial position 1, a pair of blades AA, rotating with a minimum angular velocity, separates the inlet and outlet windows 12 and 13. A pair of blades B-B begins to move (rotation) with acceleration in the direction of the arrow. In this case, a chamber 1 is formed, into which the working mixture is sucked.

 In position 2, a pair of blades AA moves to an insignificant angle, and a pair of blades B-B is in a position where it reaches maximum speed.

 In position 3, the blade of a pair A-A, moving slowly, closes the inlet 12 and from this moment the pair begins to move (rotation) with acceleration, compressing the working mixture in zone 1.

 In position 4, a pair of blades AA closes with the blades BB, compressing the working mixture, which is fed under the spark plug 14. The ignition (ignition) of the working mixture occurs.

 In position 5 in zone III, the process of combustion of the working fluid occurs, which requires the appropriate time. Blades AA and BB move from position 5 to position 6.

 At position 6, a pair of AA blades occupies a dead center position of the motion conversion mechanism. From this position, under the influence of the flywheel and gas pressure (zone III) in zone III, the expansion stroke begins. At the same time, an intake stroke occurs in zone I, a compression stroke in zone II, and exhaust gas discharge in zone IV.

 Thus, the engine completes the full cycle of a 4-stroke internal combustion engine with the simultaneous flow of all four cycles, which is the undoubted advantage of this image. Moreover, the engine makes 4 working strokes per revolution of the output shaft.

 Variants of the proposed engine in the performance of DPS, i.e., a semi-external combustion engine, were considered. However, as noted, each of these options may have equivalents or sub-options for the external combustion engine, fundamentally similar to the Stirling engine.

 The use of DPS as a basis for creating options for an external combustion engine makes it possible to use atmospheric air (open system) and hydrogen or helium (closed system) as a working fluid. In addition, the choice of the variant with the translational movement of the pistons or the RPD variant has a direct impact on the technical solution of the heating chamber. In this regard, two groups of options for an external combustion engine are considered - a group with translational movement of pistons (options V and VI) and a RPD group (options VII and VIII), both with open and closed systems.

 As shown in FIG. 11a (option V) cold air enters the compressor 1 from the atmosphere, which is then pumped into the heating chamber 2. At the same time, hot combustion products of the air-fuel mixture flow from the combustion chamber 3 to the heating chamber 2, washing the combustion chamber and discharged into the atmosphere. When air is heated in the heating chamber, as its temperature rises, so does the pressure. Heated air under pressure enters from the heating chamber 2 into the working cylinder 4, where it is used as a working fluid. From the working cylinder 4, the exhaust air enters the combustion chamber 3, where it forms a fuel-air mixture with fuel burned to heat the air in the heating chamber 2.

 In embodiment VI shown in FIG. 11b, instead of air, helium or hydrogen is used as a working medium, which determines the use of a closed system and, as a result, an exhaust gas cooler 5, and air is taken from the outside only to form a fuel-air mixture in combustion chamber 3. With this scheme, compressor 1 enters instead of air, the cooled gas (helium, hydrogen), then injected into the heating chamber 2. Heated by the combustion products of the air-fuel mixture in the combustion chamber 3, gas under pressure exits the heating chamber 2 into the working cylinder 4, where it produces work. After the transfer of a significant part of the thermal energy, the exhaust gas (working medium) is discharged from the working cylinder 4 to the cooler 5, where it is cooled and again fed to the compressor, after which the cycle repeats. The air supply to the combustion chamber 3 occurs regardless of the circulation system of the working fluid.

 Options V (Fig. 11a) and VI (Fig. 11b) are versions with the translational movement of the pistons with open and closed systems, respectively.

 Options VII (Fig. 11c) and VIII (Fig. 11d) are RPD equivalents of variants V (Fig. A) and VI (Fig. B), respectively.

Claims (4)

 1. An engine comprising a compressor with an inlet element, an external combustion chamber connected to the compressor through a first bypass element and having an ignition source configured to be switched off during operation, an expander with an exhaust element in communication with the combustion chamber through a second bypass element, and workers bodies located in the compressor and expander and kinematically connected with the output shaft, moreover, the compressor, combustion chamber and expander are equipped with individual cooling systems, and the chamber with The scuffing is performed with a volume exceeding the volume of the compressor, characterized in that the kinematic connection of the working bodies with the output shaft is made in the form of a planetary mechanism with two bevel gears, two satellites and a carrier with a gear ring, and the compressor working element is connected to one of the bevel gears, the output shaft - with another bevel gear, and the working body of the expander - with a gear ring drove.  2. The engine under item 1, characterized in that each working body is made in the form of at least one pair of blades.  3. The engine according to paragraphs. 1 and 2, characterized in that each blade is made with grooves around its perimeter and is equipped with L-shaped plate seals located in the grooves.  4. The engine according to paragraphs. 1 to 3, characterized in that each blade is made with an internal cavity filled with circulating fluid.

Images