RU2472023C2 - Operating method (versions) and arrangement of compound engine with two-phase working medium based on piston-type internal combustion engine (versions) - Google Patents

Abstract

FIELD: engines and pumps.

SUBSTANCE: basic piston-type engine is equipped with converted steam cylinder(s) instead of gas one(s), from which power is transmitted to compound engine. Combustion products are cooled in gas cylinders owing to water spraying in inlet air and their evaporation during fuel burning. Steam for operation of steam cylinder(s) is generated in steam generator as a result of feed water heating with oil heat of lubrication system and combustion products. Water drops evaporating in cylinders during fuel burning decrease thermal factor of the engine, suppress detonation and allow increasing combustion degree of petroleum and gas (with forced ignition) of engines, as well as increasing the quantity of working medium. Waste steam in cylinder(s) is condensed in condenser together with water steams and hydrocarbon steams from combustion products due to their cooling with cold water spraying.

EFFECT: reduction of hazardous emissions as a result of reduction of decomposition of combustion products, condensation of hydrocarbon steams from combustion products, and owing to content in water drops of sprayed ammonia and burning of fuel at excess air factor of more than 1.

31 cl, 19 dwg

Description

The invention relates to mechanical engineering and can be used in internal combustion engines to increase their efficiency, improve their fuel efficiency, reduce harmful emissions into the atmosphere, and increase their reliability.

A known method of operation and device piston internal combustion engine with a gas-steam working fluid (RF patent 2232913, CL F02G 5/04, F01K 23/14, F02B 47/02, 75/10).

The device includes a reciprocating internal combustion engine, a condenser, cooling devices for combustion products, water sprayers installed in the intake system and connected to the water tank by water pipes through a flow regulator and a water pump. Along the exhaust system, a steam generator, combined with an exhaust manifold, a contact condenser, a turboexpander (turbine) or a screw expansion machine, connected mechanically to an electric generator, and a moisture separator are arranged sequentially along the gas flow. Along the intake system, an air filter, a compressor, a contact air cooler and an intake manifold are arranged sequentially along the air flow. In this case, the dehumidifier tray is connected by a water supply system to the cold part of the water accumulator, the contact condenser and air cooler trays are connected by water supply to the hot part of the water accumulator, the steam generator is connected by the inlet water supply to the hot part of the water accumulator and a discharge steam line with the steam accumulator, from which the steam pipelines to the turbine or expansion screw are allotted a machine connected mechanically to a compressor.

The method of operation of this piston engine is implemented by supplying air to the cylinders of the engine, which is driven by a steam turbine. The combustion products are cooled in a steam generator, a contact condenser and a turboexpander (turbine), and the charge air in a contact air cooler. Steam is supplied to the turbine from the accumulator, which is generated by the heat of the combustion products. Water is injected into the combustion chambers from the hot part of the battery, where it is supplied from the condenser. The cold part of this battery is supplied with water released in the dehumidifier behind the turboexpander (turbine), while regulating the pH of hot water above 7 coming from the condenser into the water accumulator, adding ammonia to this water and the amount of air supplied by the compressor in accordance with the minimum allowable excess coefficient air.

The disadvantages of this known method of operation and the design of a piston engine are complexity, large dimensions and mass, as well as insufficiently high efficiency, which should be limited to the actions of the first and second laws of thermodynamics.

The closest to the claimed method of operation and the device of a combined engine with a two-phase working fluid based on a reciprocating internal combustion engine is the technical essence and the achieved result, the method of operation and the device of a combined engine with a two-phase working fluid based on a reciprocating internal combustion engine, published on the Internet website www. wipo.int/pctdb No. WO 2008/094071 dated 10/16/2008.

The basic equipment of this engine is a reciprocating internal combustion engine, a steam expansion machine or steam (s) cylinder. This expansion machine or this (s) cylinder (s) is connected (s) to the steam generator and condenser. To increase the temperature of steam at low temperatures of the combustion products, a closed circulation loop is organized through the steam generator. The basic piston engine is used with and without a traditional cooling system. In the latter case, the working fluid is cooled by spraying cold water in it with certain average droplet sizes and in the required amount. The volume of the combustion chamber in gas cylinders with forced ignition of the fuel is set so that the effective efficiency of the combined engine is greatest and there is no detonation in the cylinders due to atomization of water with the required average droplet sizes in the required amount in the inlet duct and heating it to a boiling point by the end of compression with air, as well as cooling the combustion products during fuel combustion by evaporating drops of water heated to a boiling point.

Steam for operating the steam expansion machine or feeding it into the steam cylinder (s) is obtained in the steam generator by heating the feed water with the heat of the oil of the lubrication system, water of the cooling system, as well as combustion products. In the absence of a traditional cooling system, the combustion products are cooled in cylinders by the evaporation of heated drops of water. These droplets are formed in the intake system by spraying chilled water in the radiator with outside air, and they are heated to the boiling point of the compressible and heated air during compression both in the compressor and in the cylinders. Evaporating droplets in the cylinders during fuel combustion suppress detonation and increase the compression ratio of the engine with forced ignition, as well as increase the amount of working fluid. The exhaust steam is condensed in the condenser, and water vapor and hydrocarbons from the combustion products also due to their cooling by spraying cold water, which is obtained from the condensate by cooling it in the radiator with outside air. Reduction of harmful emissions occurs mainly as a result of a decrease in the dissociation of combustion products, condensation of hydrocarbons from the combustion products on the surfaces of water droplets in the condenser and their combustion along with atomized water in the combustion chambers, and also due to the content of dissolved ammonia in the water droplets.

The main disadvantages of this prototype are the lack of reliable operation at low outdoor temperatures, the complexity of the design, and the problem of increasing effective efficiency is solved mainly in gasoline (with forced ignition) engines.

The aim of the invention is to increase the efficiency of not only a gasoline (with forced ignition) engine, but also gas and diesel engines to a level limited by the first and second laws of thermodynamics, to increase the reliability of these engines when they are used in the field of low outdoor temperatures, and also to reduce the complexity their manufacture and cost reduction.

This goal is solved by the fact that in the method of operating a combined engine with a two-phase working fluid, including supplying air to the intake manifold by a compressor, cooling it by spraying with cold water, cooling the combustion products in series in the boiler and in the contact condenser to a temperature at which the condensate released is cooled to a temperature cold water with outside air in a water-air radiator, then their expansion in a turboexpander (turbine), separation of water droplets in a contact condenser and a turbo when expanding the combustion products, supplying cold water to the contact condenser for cooling the combustion products and to the intake system for cooling the air, lowering the maximum combustion temperature due to the evaporation of water droplets when burning fuel in the combustion chambers, heat transfer from the oil of the heated water lubrication system in the water distribution device through the dividing walls of the oil-water heat exchanger, the supply of heated water to the boiler from the water distribution device and the supply of steam from the boiler to the expansion The device for obtaining useful work, regulating the gas pressure in the condenser by throttling it in front of the impeller of the turboexpander (turbine) depending on the outdoor temperature, they are carried out as a result of lowering the maximum temperature of combustion due to the evaporation of water droplets when burning fuel in the combustion chambers with air excess factors α = 1.0 ÷ 1.2 in gas and gasoline (with forced ignition) and the lowest, but greater than 1.1 in diesel (self-igniting) engines, with cat ryh organize complete combustion of the fuel in the absence of chemical incomplete combustion. Moreover, the water is fed into the air entering the compressor of a gasoline (with forced ignition) engine to reduce the temperature of this air when it is compressed in the compressor, while spraying it with such droplet sizes and in such an amount that the compression ratio of the engine increases without causing detonation, and effective efficiency becomes greatest. In addition, in a diesel (self-igniting) engine, heated water is supplied in the boiler to a boiling point at a pressure higher than the maximum combustion pressure in gas insulated gas cylinders at the beginning of expansion when burning fuel with the lowest excess air coefficient, but ≥1.1 ensuring high-quality mixture formation and complete combustion of fuel, at the same time they switch to partial loads by quantitative power control (by changing the amount of incoming air into the cylinders at a constant coefficient of ytka air). Moreover, water is sprayed in the heated air when it is compressed in the compressor and cylinders, in which there is a reliable ignition of fuel droplets in a diesel (self-igniting) engine in all modes with an allowable rigidity of its operation. In addition, they increase the effective power and pressure of the charge air due to an increase in the triggered heat drop on the impeller of the turboexpander (turbine) as a result of a decrease in the bore of the guide nozzle apparatus while maintaining the coefficient of excess air in the cylinders α = 1.0 ÷ 1.2 in gas and gasoline ( with forced ignition) and α≥1,1 in diesel (self-igniting) engines. In addition, in gas and gasoline (with forced ignition) engines, a constant or most advantageous ignition timing is established in all operating modes, and the resulting detonation is suppressed by increasing the amount of cold water sprayed in the air entering the cylinders. Moreover, the amount of feed water to the boiler or the temperature and pressure of the water (steam) are changed depending on the engine operating mode, while the combustion products in this boiler are cooled to 150 ° C. In addition, the generated steam in the boiler is fed into the gas cylinders at the beginning of expansion through the inlet steam valves and swirls at a higher compression ratio than the base engine with a pressure exceeding the gas pressure in these cylinders in such an amount and with a temperature at which detonation occurs in the combustion chambers of gas and gasoline (with forced ignition) engines, and in the cylinders of a diesel (self-igniting) engine, efficient operation with the highest efficiency is performed. Moreover, the amount of feed water to the boiler is regulated by the temperature of the exhaust gases at the outlet of this boiler, or by the pressure or temperature of the steam at the outlet of the boiler. In addition, a mixture of gasoline and water is sprayed in the air entering the cylinders, in which the requirements for the octane number of the fuel are the least, the degree of compression and the most efficient. Moreover, a mixture of water and diesel fuel is sprayed in a ratio of 0.5 ÷ 1.25 from the cyclic supply of fuel in the air entering the cylinders, in which the effective efficiency of the diesel engine (with self-ignition) becomes the highest with the permissible rigidity of its operation not exceeding that of existing engines. In addition, steam is supplied to the steam cylinder (s) through the inlet valve (s) when it is opened at the top (inner) dead center, and it (they) is closed to allow a decrease in the steam pressure in the steam generator at which the effective engine efficiency begins to decline. Moreover, with a decrease in the water level in the water and cold and hot water distribution devices, they increase the gas pressure in the condenser or add water to these tanks, and when the engine is off and the cold water temperature drops to freezing temperature, the engine starts automatically or manually. In the event that the temperature of cold water in the engine drops to freezing, water from the water and steam supply system is automatically or manually drained through a pipeline into an insulated tank with an electric pump, and when it is started, it is heated with an electric heater and refilled into this system from this tank. In addition, the goal is solved by the fact that in the method of operation of the combined engine with a two-phase working fluid, the gas pressure in the condenser is created as a result of their throttling and expansion in the throttle valve, and air is supplied to the cylinders by suction through the intake manifold, while cold water is sprayed in this air in all modes with such average droplet sizes and in such a quantity that the effective efficiency of a gasoline (with forced ignition) engine cycle becomes greatest with increased compression ratio and operating without detonation.

The goal in the device of a combined engine with a two-phase working fluid is achieved by the fact that it contains gas cylinders of the base piston engine for expanding combustion products, a boiler combined with an exhaust manifold, a condenser with a turboexpander (turbine) located in series along the exhaust gas pipeline and connected by gas communication with the cylinders of this engine, a compressor connected to the cylinders for air supply and connected to a turboexpander (turbine), as well as water distribution devices hot and cold water, interconnected with a condenser and air-water radiators by pipelines, fuel nozzles in the inlet pipes, connected by pipelines with a fuel pump, a filter and a flow regulator for supplying fuel to the cylinders, as well as heat-insulated parts of the combustion chamber [head (covers) cylinders, pistons and gas cylinders], a boiler combined with a heat-insulated exhaust manifold, nozzles and a high pressure pump with a flow regulator for supplying water to the cylinders from the boiler at the beginning asshireniya, wherein the nozzle holes are arranged nozzles to the combustion chamber walls tangentially or at an acute angle to flowing steam mixture from the nozzles organized vortical motion hot mixture in the combustion chamber to improve mixing and combustion process. In addition, the device of a combined engine with a two-phase working fluid contains gas cylinders of the base piston engine for expanding combustion products, a boiler combined with an exhaust manifold, a condenser with a turboexpander (turbine), arranged in series along the exhaust gas pipeline and connected by gas communication with the cylinders of this engine, a compressor connected to cylinders for air supply and connected to a turboexpander (turbine), as well as hot and cold water distribution devices interconnected with a condenser and a water-air radiator by pipelines, fuel nozzles in the inlet pipes, connected by pipelines with a fuel pump, filter and flow regulator for supplying fuel to the cylinders, heat-insulated parts of the combustion chamber [cylinder head (covers), pistons and gas cylinders] , a boiler combined with a thermally insulated exhaust manifold, as well as steam inlet valves with swirls in the thermally insulated cylinder head (s), connected by pipelines to the boiler to which water is supplied by a feed pump through a water flow regulator. In addition, it contains the heat-insulated steam (s) cylinder (s), the boiler is connected to the inlet valve (s) of this (s) cylinder (s) by the heat-insulated steam pipe (s), and a condenser is connected to the outlet valve (s) with a thermally insulated (s) heat conduit (s) through a non-return valve (s) to prevent the steam or gases from flowing back into the cylinder (s), while the inlet (s) the valve (s) starts (s) to open at top dead center, and closes (s) at a vapor pressure in the cylinder (s) at which the effective engine efficiency with ANOVA largest. In addition, the device of this engine contains a thermally insulated tank with high resistance to heat transfer from the walls, air and steam valves, a filler neck, which is connected through an electric pump and shut-off valve to the water and steam supply system, this electric pump is connected to the control panel, and this tank can be disconnected from this system and transferred to a warm storage room, as well as heating water in it to heat the engine before starting. In addition, the device of the combined engine contains heat-insulated water distribution devices for hot and cold water, a condenser, a boiler, a water-air radiator, connecting water pipes and shut-off and control valves, the designs of which allow reliable water discharge from a two-phase water and steam supply system manually or automatically when the temperature drops cold water to freezing temperature, as well as to fill this system with water before starting the engine. It also contains a hot water distribution device divided by perforated partitions into three parts, in the middle part of which there is a water-oil heat exchanger, condensate supply pipes from the condenser and heated feed water to the boiler, and condensate supply pipes from the condenser and mixed water discharge to the bottom are located cold water distribution device, while the cross section of the hot water distribution device is selected so that the speed of movement of water in the perforation The openings of the lower baffle were higher than the hydrocarbon float rate (0.2–0.3 m / s) so that they could be transferred to a cold water distribution device for the purpose of feeding them into the air entering the cylinders. Further, this engine contains nozzles in the inlet nozzles connected to the water supply pipes for supplying cold water from the water distribution device on which the pump and the flow regulator are installed, and also has a carburetor connected to the intake manifold and the fuel pump by a pipe on which the fuel filter is located, while nozzle nozzles are sized so that they form droplets that turn into a working fluid when fuel is burned, and the pump is designed so that it delivers water Naturally with a flow regulator 0.5 ÷ 1.25 from cyclic fuel supply at the highest compression ratio and operation without detonation. The design of this engine contains a carburetor connected to the intake manifold, a fuel line for supplying fuel with a fuel pump through a fuel filter and a fuel flow regulator, as well as a cold water pipe for supplying it with a water pump through a water flow regulator, while mixing fuel with water is provided in the float chamber, and the dimensions of the carburetor atomizer and its diffuser have certain dimensions to obtain the required droplet size in air, at which the effective engine efficiency becomes tsya greatest at the highest compression ratio operation without knocking.

In addition, the engine contains a fuel and water mixer, which is connected to the cold water distribution device by two water pipes. The first of them has a water pump, and the second has a flow regulator. It is also still connected to the fuel pump by a fuel line, on which the fuel filter and the fuel consumption regulator are located, moreover, this mixer is still connected by pipelines through the pump and the flow regulator to the nozzles of a gasoline (forced ignition) engine located on the inlet manifold nozzles, or it is connected by a pipeline through a high-pressure pump, a flow regulator and a pressure reducing valve with nozzles of a diesel (self-igniting) engine installed in the heat-insulated head (s) cylinder), while the sizes of the nozzle nozzles are selected so that the droplets formed from the mixture of water and fuel turn into a working fluid when the fuel is burned, and the design of the pump with a flow regulator can provide water supply 0.5 ÷ 1.25 from the cyclic fuel supply at the highest compression ratio and work without detonation. In the device of this engine, a closed circulation circuit is formed under the action of the feed pump through the boiler and the first locking device associated with the logical telecommunication device. For normal operation, a second shut-off device and a valve are installed on the steam line connecting the boiler to the inlet (s) of the steam valve (s) of the steam cylinder (s), to communicate this (their) steam cylinder (s) o) with the atmosphere to reduce the operation of the pumping strokes of the disconnected (s) steam (s) cylinder (s), which are connected by telecommunications with the logical device. In the cylindrical engine devices used, the nozzles for supplying and removing liquids of a cylindrical shape are located tangentially to the inner cylindrical surfaces of the tanks for mixing heat carriers with different temperatures, as well as increasing the heat transfer coefficient through the dividing walls of the oil-water heat exchanger, while the nozzles for water drain are located vertically in the center of these devices and have perforated holes for the intake of mixed water, and they exit through the side, top or bottom e wall outwardly. In addition, this engine may contain a water pump and nozzles for adding water to the air in the inlet nozzles of gas and gasoline (with forced ignition) engines and a water pump and nozzles for supplying water to the insulated cylinders of a diesel (self-igniting) engine and spraying it there into air, while the dimensions of the pumps and atomizers of nozzles are such that water is sprayed in an amount and with such average droplet sizes, which upon compression becomes a heat carrier, and upon combustion, eraticates into the working fluid. To control the operation of the engine in manual and automatic modes, there is a control panel that is connected to temperature, pressure, detonation and coolant flow sensors, as well as logical and executive devices. In addition, the device is a combined piston gas or gasoline (with forced ignition) and diesel (self-igniting) engine, containing a reciprocating internal combustion engine, a capacitor located in series along the exhaust gas pipeline and connected by gas communication with the cylinders of this engine, as well as an intake manifold, characterized the fact that the exhaust gas pipeline is equipped with a throttle valve for regulating the gas pressure in the condenser, in which water sprayers are installed for lowering the temperature of the combustion products, in order to supply fuel in a gasoline (forced ignition) engine, nozzles are installed in the inlet pipes, and in a diesel (self-igniting) cylinder they are placed in the heat-insulated cylinder head (s), the air is supplied by suction through the intake manifold and regulate the amount of air supplied to the cylinders with a throttle valve installed in the inlet air duct at various operating modes, and the steam is supplied through the steam valves through the steam line the boiler, the steam pressure in the steam line feed pump is provided, and its amount determined by a flow regulator.

при различной степени сжатия ε, одинаковых температуре начала сжатия Т_a=293 К и одинаковых коэффициентах избытка воздуха α=1 и наполнения η_ν=1:1-ε=8; 2-ε=10; 3-ε=12; 4-ε=14. На фиг.10 показано изменение давления Р_с и температуры Т_c в конце сжатия в зависимости от доли количества впрыснутой воды, от количества поданного топлива,

при различной степени сжатия ε, одинаковой температуре начала сжатия Т_a=293 К и одинаковых коэффициентах избытка воздуха α=1 и наполнения η_ν=1:1-ε=8; 2-ε=10; 3-ε=12; 4-ε=14. На фиг.11 изображено изменение максимальных температуры T_z и давления Р_z, а также КПД теоретического цикла η_i в зависимости от количества впрыснутой воды, от количества поданного топлива,

при различной степени сжатия ε, одинаковой температуре начала сжатия T_a=293 К и одинаковых коэффициентах избытка воздуха α=1 и наполнения η_v=1: 1-ε=8; 2-ε=10; 3-ε=12; 4-ε=14. I - изменение η_i при T_z=const и увеличениях ε и

II - изменение Р_z при T_z=const и возрастаниях ε и

На фиг.12 приведены кривые изменения среднего эффективного давления Р_е, температуры горючей смеси в конце впуска Т_а, коэффициентов наполнения η_v и остаточных газов γ, а также температуры остаточных газов Т_r, в зависимости от доли впрыснутой воды, от количества поданного топлива,

при различных коэффициентах избытка воздуха α: 1-α=0,93; 2-α=1,0; 3-α=1,08;

- точки по экспериментальным данным,

- расчетные кривые. На фиг.13 показано изменение среднего эффективного расхода топлива, эффективного η_е и индикаторного η_i КПД, доли теплоты, превращенной в эффективную работу, q_e, давления конца сжатия Р_с в зависимости от доли впрыснутой воды, от количества поданного топлива,

при различных коэффициентах избытка воздуха α: 1-α=0,93; 2-α=1,0; 3-α=1,08;

- точки по экспериментальным данным,

- расчетные кривые. На фиг.14 изображено понижение максимальной температуры газов на единицу повышения максимального давления газов для устранения детонации ΔТ_z/ΔР_z в зависимости от максимальной температуры газов Т_z. На фиг.15 показано понижение максимальной температуры газов на единицу повышения максимального давления газов для устранения детонации ΔТ_z/ΔР_z в зависимости от доли количества теплоты, теряемой в систему охлаждения q_ox. На фиг.16 изображено понижение максимальной температуры газов на единицу повышения максимального давления газов для устранения детонации ΔT_z/ΔР_z в зависимости от доли количества теплоты, теряемой с выпускными газами q_вып.Figure 1, 2, 3, 4, 5, 6, and 7 show embodiments of the working process of a combined engine with a two-phase working fluid. Figure 1 shows a diagram of the supply of fuel to the intake air by nozzles located in the inlet pipes; water supply to the compressor intake air by a nozzle for quasi-isothermal compression of this air; as well as the supply of heated water in the boiler to the nozzles located in the cylinder head (s). This diagram also shows a condenser, a turboexpander (turbine), as well as water distribution devices for cold and hot water, their location and interaction. The diagram of figure 2 shows the operation diagram of this combined engine without a turboexpander (turbine) and compressor, quantitative regulation of its power and the amount of condensate formed. In addition, the steam supply from the boiler to the gas cylinders through the inlet steam valves during the expansion of gases in these cylinders is shown. Figure 3 shows the conversion of one of the gas cylinders into steam (for steam operation). In this figure, water distribution devices for cold and hot water are located, as well as a thermally insulated container for draining water from the water and steam supply system when its temperature drops to freezing temperature. This diagram shows the utilization of the heat of the used oil in a hot water distribution device, as well as the condensate cooling with outside air in a cold water distribution device. Figure 4 shows the atomization of fuel by a carburetor in the intake manifold, as well as the supply of cold water from the water distribution device to the nozzles installed in the inlet pipes. Figure 5 shows the mixing of fuel and water in a carburetor and the atomization of the fuel-water mixture by this carburetor in the intake manifold. Figure 6 shows the mixing of fuel and water in a special mixer and the supply of this mixture to the nozzles of gasoline (forced ignition) and diesel (self-igniting) engines. In Fig.7, the coolant is circulated through the boiler at a reduced temperature of the exhaust gases to increase the temperature of steam or water. On Fig shows the supply of coolant to the tank and its removal from it, where there is a mixing of liquids with different temperatures. Figure 9 shows the curves of changes in the temperature of water droplets at the end of compression t _water.k , the fraction of heat transferred to the droplets of water during compression, on the heat of combustion of the fuel, as well as the average pressure of the theoretical cycle P _i depending on the proportion of the amount of water dispensed, from the amount of fuel supplied

at different degrees of compression ε, the same temperature of the beginning of compression Т _a = 293 K and the same coefficients of excess air α = 1 and filling η _ν = 1: 1-ε = 8; 2-ε = 10; 3-ε = 12; 4-ε = 14. Figure 10 shows the change in pressure P _c and temperature T _c at the end of compression, depending on the proportion of the amount of water injected, on the amount of fuel supplied,

at different degrees of compression ε, the same temperature of the beginning of compression Т _a = 293 K and the same coefficients of excess air α = 1 and filling η _ν = 1: 1-ε = 8; 2-ε = 10; 3-ε = 12; 4-ε = 14. 11 shows the change in the maximum temperature T _z and pressure P _z , as well as the efficiency of the theoretical cycle η _i depending on the amount of water injected, on the amount of fuel supplied,

at different compression ratios ε, equal compression start temperature T _a = 293 K and the same excess air ratio α = 1, and filling η _v = 1: 1-ε = 8; 2-ε = 10; 3-ε = 12; 4-ε = 14. I is the change in η _i at T _z = const and increases ε and

II - change in P _z at T _z = const and increasing ε and

On Fig curves of changes in average effective pressure P _e , temperature of the combustible mixture at the end of the inlet T _a , filling ratios η _v and residual gases γ, as well as the temperature of the residual gases T _r , depending on the proportion of injected water, on the amount of fuel supplied ,

at various coefficients of excess air α: 1-α = 0.93; 2-α = 1.0; 3-α = 1.08;

- points according to experimental data,

- calculated curves. On Fig shows the change in average effective fuel consumption, effective η _e and indicator η _i efficiency, the proportion of heat converted into efficient operation, q _e , pressure of the end of compression P _with depending on the proportion of injected water, on the amount of fuel supplied,

at various coefficients of excess air α: 1-α = 0.93; 2-α = 1.0; 3-α = 1.08;

- points according to experimental data,

- calculated curves. Fig. 14 shows a decrease in the maximum gas temperature per unit of increase in the maximum gas pressure to eliminate detonation ΔT _z / ΔP _z depending on the maximum gas temperature T _z . On Fig shows a decrease in the maximum gas temperature per unit increase in the maximum gas pressure to eliminate detonation ΔT _z / ΔP _z depending on the fraction of the amount of heat lost in the cooling system q _ox . In Fig.16 depicts a decrease in the maximum gas temperature per unit increase in the maximum gas pressure to eliminate detonation ΔT _z / ΔP _z depending on the fraction of the amount of heat lost with the exhaust gases q _issue .

In order to prove the industrial applicability of this invention, before the preparation of the application for this invention, the authors carried out research work on the influence of the quality and quantity of atomized water in the intake air on the parameters and performance of gasoline engines. The reason for increasing the efficiency of gasoline engines when spraying water in the intake air was established, despite the decrease in the maximum combustion temperature. The essence of this reason is that when water is sprayed in the intake air with a certain size and quantity, they turn into a coolant when compressed, heating to a boiling point almost without evaporation, and at the same time reduce the temperature and pressure of the working fluid. As a result, the compression work in the cylinder decreases. Further, when isochoric supply of heat to the working fluid, these drops begin to quickly evaporate and turn into steam. As a result, the number of working fluid increases, and consequently, the amount of work performed during expansion. An increase in the expansion work occurs not only due to an increase in the quantity of the working fluid, but also changes in its composition, despite a decrease in the maximum cycle temperature. The composition of the working fluid affects the amount of work done through the degree of expansion.

To establish this reason, a mathematical model of thermal calculation of the cycle of gasoline engines was developed taking into account the addition of water to the combustible mixture, which is described below.

The mathematical model of thermal calculation is based on the method developed by professors of MSTU. N.E.Bauman V.I. Grinevetsky and E.K. Masing.

This calculation according to the mathematical model begins with determining the amount of the combustible mixture entering the cylinder, as well as the amount and composition of the combustion products formed by compressing 1 kg of fuel.

The mass of the combustible mixture, taking into account the supply of water vapor to the inlet pipe with the known coefficient of excess air α and theoretically necessary amount of air L _o for burning 1 kg of fuel will be determined

- количество водяного пара, поданного на 1 кг топлива, кг/кг;Where:

- the amount of water vapor supplied per 1 kg of fuel, kg / kg;

L _o - theoretically necessary amount of air for burning 1 kg of fuel, kg / kg;

α is the coefficient of excess air.

The mass of the combustible mixture when water is supplied to the intake manifold and sprayed in air is determined by burning 1 kg of fuel

In the method of calculating Grinevitsky-Masing pressures at the end of the filling process P _a and at the end of the production process P _r are set according to the data of existing engines. Here they are calculated because they change when water vapor is supplied to the intake manifold. Intake pressure

where: p _{o (k))} is the pressure of the combustible mixture in the inlet pipe, Pa;

ρ _K.s. - the average density of the combustible mixture in the intake organs, kg / m ³ ;

ζ _VP = 2.5 ÷ 4 - average for the filling process, the total coefficient of hydraulic losses in the intake organs, calculated by the average speed of movement of the combustible mixture;

ω _avg - average speed of the combustible mixture in the intake organs, m / s.

The value of ω _{avg is} determined by the equation

where: F _VP.cp is the average passage section of the intake organs in m ² ;

G _m - second fuel consumption in kg / s;

i _pat.vp - the number of inlet pipes.

The value of the second fuel consumption is determined

- удельный эффективный расход топлива для первого приближения в кг/(кВт·ч);Where:

- specific effective fuel consumption for the first approximation in kg / (kW · h);

- эффективная мощность двигателя для первого приближения в кВт.

- effective engine power for a first approximation in kW.

The average pressure in the exhaust process is determined in the same way as P _a

where: p _p - gas pressure in the exhaust pipe, Pa;

ρ _R.SR - the average density of gases in the exhaust organs, kg / m ³ ;

ζ _vyp is the average total coefficient of hydraulic losses in the exhaust organs for the exhaust process, calculated by the average gas velocity;

ω _SR.VYP - average velocity of exhaust gases in the exhaust organs, m / s.

The value of ω _{SR.VYP is} determined by the equation

where: G _{PR. SG} - mass of combustion products, kg / kg;

i _PAT.VYP - the number of exhaust pipes;

F _{Vyp. SR} - the average passage section of the exhaust organs in m ² .

Knowing P _a and P allows you to determine the average pressure of the pumping _pass P _n . The value of P _{nx is} necessary to determine the average actual indicator pressure P _i∂ .

We write the heat balance equation to determine the temperature of the combustible mixture at the end of filling T _a , taking into account the water supply to the inlet pipe and its heating. Moreover, it is accepted that gasoline is heated to T _a and the whole evaporates. This equation for 1 kg of fuel is written:

where: T _{O (K)} is the outdoor temperature after the compressor in the absence of an air cooler or after an air cooler in its presence. In the absence of a compressor and an air cooler, the outdoor temperature, K;

T _BEN - temperature of gasoline supplied to the mixture formation, K;

T _{r '} is the temperature of the residual gases, which expand from r _r to p _a , K;

,

,

- теплоемкости воздуха, бензина и продуктов сгорания при

;

,

,

- heat capacity of air, gasoline and combustion products at

;

, ,

- теплоемкость воздуха, бензина и продуктов сгорания при V=const и T_о(к), Т_БЕН, Т_r', соответственно,

;

, ,

- the heat capacity of air, gasoline and combustion products at V = const and T _{about (k)} , T _BEN , T _{r '} , respectively,

;

G _cart = α L _O - the actual amount of air required for combustion of 1 kg of fuel at α, kg / kg;

G _r - the amount of residual gases in the cylinders with the piston in. mt formed during the reduction of 1 kg of fuel, kg / kg;

- теплота парообразования 1 кг бензина;

- heat of vaporization of 1 kg of gasoline;

N _u is the calorific value of fuel, J / kg;

ΔН _u - heat loss from theoretical incompleteness of combustion, J / kg;

υ _a is the fraction of the calorific value of 1 kg of fuel, which characterizes heat transfer during the filling period between air, cylinder walls and water drops.

From equation (8) we determine the temperature of the mixture at the end of the filling

Equation (9) for determining T _a in K when water is supplied to the inlet pipe is written

и

- теплоемкости водяного пара при V=const, и температурах Т_ВОД и Т_а,

;Where:

and

- the heat capacity of water vapor at V = const, and temperatures T _WATER and T _a ,

;

υ _a is the fraction of the calorific value of 1 kg of fuel, which characterizes the heat transfer during the filling period between the combustible mixture and the cylinder walls.

In the Grinevecki-Masing method of calculation, the value of T _r for the thermal calculation is specified in the source data. Here it is taken for a first approximation, and then refined after calculating the release process by calculating P _r according to equation 6.

The filling factor of the combustible mixture of cylinders will be determined by the ratio of the actual amount of the combustible mixture entering the cylinder, G _factor to theoretical G _theory .

When water is supplied to the inlet pipe, the equation for determining the G _ph , taking into account the decrease in the working volume of the cylinder with water drops,

- уменьшение рабочего объема цилиндра каплями воды, м³;Where:

- reducing the working volume of the cylinder with water drops, m ³ ;

- количество воды, впрыснутой во впускной трубопровод, кг;

- the amount of water injected into the intake pipe, kg;

ζ ₁ - recharge coefficient.

The value of ζ ₁ can be determined by the known P _a and the gas pressure at the moment of closing the inlet opening P _s.k. , as well as the total volume of the cylinder V _a and its volume at the time of closing the inlet bodies V _Z.K.

Theoretical amount of combustible mixture that can enter the cylinder when water is supplied to the inlet pipe, taking into account the reduction in the cylinder’s working volume with water drops

The equation for the compression process of a combustible mixture based on the first law of thermodynamics is written

where: Q _C - the amount of heat given off by the combustible mixture to the walls of the cylinder and drops of water during compression, Q _C = υ _С (Н _u -ΔH _u );

υ _C is the fraction of heat from the heat of combustion of the fuel, characterizing heat transfer during the compression between the walls of the cylinder, drops of water and air, J;

U _C and U _a - internal energy of the compressible working fluid at the beginning and end of compression, J;

,

,

where: L _ac - the work of polytropic compression, J;

- давление в конце сжатия для первого приближения, Па.

- pressure at the end of compression for the first approximation, Pa.

We substitute the values of the terms in equation (13), we obtain

from here

- цикловая подача топлива, кг/цикл.Where:

- cyclic fuel supply, kg / cycle.

Another equation for determining T _C will be

относительно показателя политропы сжатия n_C. После определения n_C находим более точное значение давления конца сжатияEquations 15 and 16 are solved numerically with the known

relative to the polytropic compression index n _C. After determining n _C we find a more accurate value of the pressure of the end of compression

;Where:

;

.

.

подставляем Р_C и вновь решаем уравнения 15 и 16 численным методом для определения более точного значения n_C.Then in equation 15 instead

we substitute P _C and again solve equations 15 and 16 by a numerical method to determine a more accurate value of n _C.

The work spent on the compression of the working fluid in the cylinder,

We write the equation of fuel combustion for a cycle with isochoric heat input and its removal to evaporate water droplets to determine the maximum theoretical combustion temperature T _Z

where: υ _Z - coefficient taking into account the amount of heat received by the working fluid during isochoric combustion of the fuel (heat utilization factor in bhp);

ΔH _u = 6100 (1-α) is the heat loss from chemical incompleteness of fuel combustion at α <1, kJ / kg.

The heat spent on the vaporization of water and heating its vapors to T _Z will be determined

where: r _water - the heat of vaporization of water is determined according to the table depending on the average gas pressure during the combustion period, J / kg;

T _water.k — temperature of water droplets heated to boiling point (saturation), K;

υ _IP - coefficient taking into account the proportion of evaporated water during heat supply (fuel combustion); υ _IS = 1 during the evaporation of all water near v.m.

Equations 19 and 20 are solved numerically with respect to T _Z.

Theoretical maximum gas (combustion) pressure during heat supply and water evaporation

Actual maximum combustion pressure is

where: φ _P = 0.7 ÷ 0.9 is the pressure reduction coefficient.

The heat balance equation for the expansion process of the working fluid is written

where: Q _ZB - change in the enthalpy of the working fluid (combustion products) during its (their) expansion, J;

U _Z and U _B - the internal energy of the working fluid (combustion products) at the beginning and end of expansion, J:

L _ZB - work is done by the working fluid (combustion products) during its (their) expansion, J.

The value of L _{Zb is} determined by the equation

We substitute the value of the terms in equation (23), which is written for the cyclic fuel supply

Solving this equation with respect to T _B , we obtain in K

where: υ _b is the fraction of the calorific value corresponding to mechanical work and the increase in internal energy in the entire area of combustion and expansion (coefficient of heat utilization at the end of expansion at the beginning of production).

Another equation for T _b will be written

Solving equations 26 and 27 by a numerical method, we find the values of the exponent polytropic exponent n _p .

для определения Т_r.With a known n _{p the} pressure in the cylinder at the end of the expansion P _b will be determined by the expression described in the method of Grinevetsky-Masing. Send

to determine T _r .

и действительное

средние индикаторные давления цикла рассчитываются по выражениям, изложенным в методике Гриневецкого-Мазинга.Theoretical

and valid

average indicator cycle pressures are calculated according to the expressions described in the method of Grinevetsky-Masing.

Indicator efficiency of the piston engine when steam is supplied to the intake manifold

Indicator efficiency of the piston engine when spraying water in the intake manifold

The effective efficiency of the piston engine η _e , specific indicator g _i and effective g _e fuel consumption, indicator N _i and effective N _e power are determined by the expressions described in the method of Grinevetsky-Masing.

и

в формуле (5).After determining N _e and g _e, their values are sent

and

in the formula (5).

Heat balance equation

where: Q ₁ = H _u · G _t.cyc - heat of combustion of spent fuel, J;

Q _ZB - heat equivalent to the expansion work, J;

Q _SJ - heat equivalent to the work of compression, J;

Q _TP = P _FUR · V _h - heat equivalent to the work of friction, J;

- теплота, теряемая с выпускными газами, Дж;

- heat lost with exhaust gases, J;

Q _X.N = ΔH _u · G _t.cyc - heat lost from chemical incompleteness of fuel combustion, J;

Q _ZB -Q _SJ = Q _e = P _e V _h - heat equivalent to perfect work, J.

In the theoretical cycle, instead of Q _{V.G., there} will be heat removed from the working fluid,

where: ℓ _KOH - heat of condensation of water vapor, J / kg;

- цикловое количество поданной воды, кг.

- cyclic amount of water supplied, kg.

The heat loss to the cooling medium Q _ox for the adiabatic engine will be zero, and for engines with a cooling system, they will be determined from the heat balance equation 30

If we express the components of the fuel balance as a percentage of the heat of the spent fuel, then equations 30 and 32 are written:

For a real adiabatic engine, Q _ОХ and q _OX = 0, and a thermodynamic cycle with isochoric supply and removal of heat Q _TP , Q _XH and Q _OX , as well as q _ТР , q _Х.H and q _OX = 0. For these engines and cycles, the correctness of the thermal calculations performed is checked by observing the heat balance equations 30 and 33.

To calculate the parameters and indicators of the cycle of a gasoline engine with polytropic compression of the combustible mixture and the expansion of the combustion products, it is necessary to set the values of the coefficients ζ _VP , ζ _SUP , υ _а , υ _с , υ _Z , υ _b , ζ ₁ and φ _p . These coefficients can be determined for existing engines using the removed and processed indicator diagrams and experimentally determined P _e , α, η _V and members of the heat balance.

When performing calculations for a real adiabatic gasoline engine, the coefficients υ _a and υ _C should be zero, and υ _b and υ _Z = 1.

If we perform a thermal calculation of the theoretical cycle with isochoric supply and removal of heat, then the coefficients υ _a and υ _C = 0, a υ _Z , υ _b , ζ ₁ and φ _p = 1, while other coefficients that determine the gas exchange process, as well as pressure gases in the exhaust gas P _p , it is necessary to set such that the coefficients of residual γ → 0 and filling η _V → 1.

When water is added to the combustible mixture, the pressure P _C and the temperature T _C at the end of compression can significantly decrease due to the occurrence of a heat exchange process between the water droplets and the working fluid. In this case, the coefficient υ _c for the theoretical cycle will not be zero. Its value will be determined in fractions of the heat of combustion of fuel N _u according to the equation

- начальная температура впрыскиваемой воды, К;Where:

- the initial temperature of the injected water, K;

T _VOD.K - the boiling point of water drops, which is determined by the pressure of the end of compression and the tables, K;

- доля воды от количества поданного топлива, кг/кг.

- the proportion of water from the amount of fuel supplied, kg / kg.

The values of P and T _{as well} may be adjusted to the initial values of the above coefficients _VI ζ, ζ and ζ _{SAT 1} and P _p so that γ → 0 and η _V → 1.

. При добавлении воды наибольшие значения эффективного КПД η_е и среднего эффективного давления Р_е получаются при

, значение υ_a=-0,002÷0,005, a υ_c=0,015÷0,025 в зависимости от коэффициента избытка воздуха α. До

υ_z и υ_в почти не изменяются, а при

начинает уменьшаться из-за снижения максимальной температуры сгорания Т_z. Увеличение η_e происходит из-за уменьшения работы сжатия и увеличения количества рабочего тела, а повышение Р_е благодаря увеличению коэффициента наполнения. С увеличением ε υ_z не изменяется, υ_c и υ_в растут, η_e быстро повышается. При увеличении ω υ_а изменяется незначительно, υ_c снижается, a υ_z и υ_в растут и достигают наибольших значений при ω=3500 об/мин. Изменение υ_z по нагрузочной характеристике происходит из-за изменения α при υ_a и υ_с≈const. При работе двигателя температура выпускных газов Т_р≥1250 К, что позволяет утилизировать их теплоту для превращения ее в работу.Using the mathematical model of thermal calculation and experimental data of the traditional MZMA-407 engine, it was found that the fraction of heat from the heat of combustion supplied to the combustible mixture during the intake process, υ _a , and the heat removed from it during compression, υ _c , as well as the heat utilization coefficients m.t. υ _z and at the end of expansion υ _b depending on the crankshaft rotation frequency ω and engine load Ne, compression ratio ε and the amount of water added to the combustible mixture

. When water is added, the highest values of effective efficiency η _e and average effective pressure P _e are obtained at

, the value of υ _a = -0.002 ÷ 0.005, a υ _c = 0.015 ÷ 0.025, depending on the coefficient of excess air α. Before

υ _z and υ _в almost do not change, and when

begins to decrease due to a decrease in the maximum combustion temperature T _z . An increase in η _e occurs due to a decrease in the work of compression and an increase in the quantity of the working fluid, and an increase in P _e due to an increase in the filling ratio. With an increase in ε, υ _z does not change, υ _c and υ _в grow, η _e rises rapidly. With an increase in ω υ _{a, it} changes slightly, υ _c decreases, a υ _z and υ _c increase and reach their maximum values at ω = 3500 rpm. The change in υ _z according to the load characteristic occurs due to a change in α for υ _a and υ _with ≈const. When the engine is running, the temperature of the exhaust gases is T _p ≥1250 K, which makes it possible to utilize their heat to turn it into work.

The nature of the changes in the coefficients of heat utilization in m.t. υ _z and at the end of the expansion υ _b , depending on the crankshaft rotation speed, they remain the same as those of the traditional MZMA-407 gasoline engine. The increase in the absolute values of υ _z and υ _b for these engines occurred due to the larger cylinder volume than the traditional MZMA-407 engine. An increase in the degree of compression in engines of the ZMZ-406.10 family affected an increase in effective efficiency through an increase in thermal efficiency, while no increase in υ _z and υ _b occurred. The inertial and gas-turbine pressurization implemented in these engines affected, first of all, their filling with a combustible mixture, and through it to increase the effective power and improve the specific effective fuel consumption. At the same time, their indicator efficiency almost did not change, and an increase in the maximum combustion pressure did not lead to detonation when operating on fuel with the same octane number due to a decrease in the maximum combustion temperature. The temperature of the exhaust gases turned out to be no lower than that of the traditional MZMA-407 gasoline engines for all external speed characteristics.

The design of these engines turned out to be acceptable for adding water to the combustible mixture in order to increase their compression ratio and eliminate the cooling system, and the temperature of the exhaust gases is sufficient to effectively convert their heat into work.

; 0,125; 0,25; 0,375; 0,5; 0,75; 1,0 и 1.25, а также сохранении коэффициентов избытка воздуха α=1 и наполнения η_V=1, а также Т_а=293 К. При этом было принято, что размеры капель воды при распыливании получаются такими, что они к концу сжатия нагреваются до температуры кипения t_ВОД.К без испарения.To conduct theoretical and theoretical studies of the effect of water supply at the beginning of compression on the parameters and parameters of theoretical cycles, calculations were performed using a mathematical model with compression ratios ε = 8, 10, 12, and 14, fractions of the amount of injected water from the amount of fuel supplied to supply heat

; 0.125; 0.25; 0.375; 0.5; 0.75; 1.0 and 1.25, as well as maintaining the coefficients of excess air α = 1 and filling η _V = 1, as well as T _a = 293 K. It was assumed that the size of water droplets during spraying is such that they are heated by the end of compression to boiling point t _WATER . _K without evaporation.

температура Т_C и давление Р_C в конце сжатия уменьшаются при всех степенях сжатия ε (фиг.10), t_ВОД.К снижается, a υ_С увеличивается (фиг.9) из-за преобладающего влияния увеличения

.Figure 9 shows that the higher the compression ratio ε, the higher the boiling point of water t _{WATER. K} due to the higher pressure of the end of compression P _C , and therefore, the higher the proportion of heat from the amount of heat input to heat the water during compression υ _C. With an increase in the amount of water supplied

the temperature T _C and pressure P _C at the end of compression decrease at all compression ratios ε (Fig. 10), t _{WATER. K} decreases, a υ _С increases (Fig. 9) due to the prevailing influence of the increase

.

максимальная температура цикла T_Z при α=1 снижается значительно благодаря увеличению количества воды, превращаемой в пар при V=const, а также снижению температуры конца сжатия Т_C. Для более высоких степеней сжатия получаются большие максимальные давления цикла при одинаковом количестве

из-за более высоких давлений конца сжатия Р_C. С увеличением

максимальное давление цикла P_Z не снижается, а даже немного увеличивается, что связано с увеличением количества рабочего тела в связи с превращением воды в пар при V=const.The maximum cycle temperature T _Z for higher compression ratios ε has higher values (11) due to the higher temperature of the compression end T _C. With an increase in the amount of water supplied

the maximum cycle temperature T _Z at α = 1 decreases significantly due to an increase in the amount of water converted to steam at V = const, as well as a decrease in the temperature of the end of compression T _C. For higher compression ratios, large maximum cycle pressures are obtained with the same amount

due to the higher compression end pressures P _C. With increasing

the maximum pressure of the cycle P _Z does not decrease, but even slightly increases, which is associated with an increase in the quantity of the working fluid in connection with the conversion of water into steam at V = const.

определяется изменениями доли потерь теплоты на сжатие q_СЖ, давления конца расширения Р_b и максимального давления цикла Р_Z, а также показателем политропы расширения n_р в зависимости от

. Для больших ε получаются и большие значения Р_i и η_i (фиг.9 и 11) независимо от количества поданной воды

. Оптимальные значения достигаются при

. При высоких степенях сжатия эти показатели располагаются ближе к 0,375, а при более низких - к 0,5. Таким образом, с увеличением количества подаваемой воды

и уменьшением максимальной температуры цикла T_Z значения η_i и Р_i не только снижаются, а даже увеличиваются и имеют оптимальное значение при

в зависимости от степени сжатия ε. Затем они начинают уменьшаться по мере увеличения

. На увеличение η_i и Р_i при небольших

оказывает превалирующее влияние увеличение количества рабочего тела по сравнению с уменьшением T_Z. Уменьшение η_i и Р_i при больших

происходит из-за преобладающего влияния на них снижения T_Z по сравнению с увеличением количества рабочего тела. При одинаковой T_Z=const значения η_i и Р_i получаются наиболее высокими при более высоких степенях сжатия ε (фиг.11), причем максимальное давление цикла P_Z становится тоже наибольшим. Следует отметить еще одно положительное свойство рабочего тела, возникающее при подаче воды, это быстрое линейное уменьшение T_Z с увеличением

и пологое изменение кривых η_i и

, что может обеспечить устойчивую работу двигателя.Change in efficiency η _i and average pressure of the theoretical cycle P _i depending on the amount of water supplied

is determined by changes in the fraction of heat loss in compression q _SJ , the pressure of the end of expansion P _b and the maximum cycle pressure P _Z , as well as the polytropic expansion index n _p depending on

. For large ε, large values of P _i and η _i are obtained (Figs. 9 and 11) regardless of the amount of water supplied

. Optimum values are achieved with

. At high degrees of compression, these figures are closer to 0.375, and at lower levels to 0.5. Thus, with an increase in the amount of water supplied

and by decreasing the maximum cycle temperature T _{Z, the} values of η _i and P _i not only decrease, but even increase and have an optimal value for

depending on the degree of compression ε. Then they begin to decrease as they increase.

. To increase η _i and P _i at small

exerts a predominant increase in the number of working fluid compared with a decrease in T _Z. The decrease in η _i and P _i at large

occurs due to the predominant influence on them of a decrease in T _Z compared to an increase in the number of working fluid. For the same T _Z = const, the values of η _i and P _i are the highest at higher compression ratios ε (Fig. 11), and the maximum cycle pressure P _Z also becomes the largest. It should be noted one more positive property of the working fluid that occurs when water is supplied, this is a rapid linear decrease in T _Z with increasing

and a gentle change in the curves η _i and

that can ensure stable operation of the engine.

изменяется несущественно, только максимумы КПД η_i и среднего давления P_i теоретических циклов с понижением Р_а и увеличением Т_а смещаются в сторону меньших

, а с увеличением Р_а и уменьшением Т_а в сторону больших

. При ε=12 они получаются смещенными в сторону меньших

, а при ε=8 в сторону больших

. Заметного влияния Т_а и Р_а на оптимальное количество

не предвиделось потому, что они сами не влияют на КПД теоретического цикла при их изменении в широких пределах. Существенное влияние на среднее давление теоретического цикла Р_i и на его КПД оказывает степень сжатия ε, поэтому на фиг.9 и 11 приведены подробно влияние

на P_i и η_i при различных ε=8÷14.In order to make sure that the detected properties of the working fluid in cycles with isochoric supply and removal of heat that occur during water supply will be preserved at other pressures P _a and temperatures T _{a of the} beginning of compression, similar calculations were performed at P _a = (0, 6 and 1.8) · 10 ⁵ Pa and compression ratios ε = 8 and 12, as well as at Т _а = 353 and 413 К and ε = 8 and 12, that is, at the values of Р _а and Т _а , which are observed for existing gasoline engines. It was found that the nature of the curves η _i and

changes insignificantly, only the maxima of the efficiency η _i and the average pressure P _{i of the} theoretical cycles with a decrease in P _a and an increase in T _a shift to smaller

, and with an increase in P _a and a decrease in T _a towards large

. At ε = 12, they turn out to be biased towards smaller

, and for ε = 8 towards large

. Noticeable effect of T _a and P _a on the optimal amount

it was not foreseen because they themselves do not affect the efficiency of the theoretical cycle when they change over a wide range. A significant effect on the average pressure of the theoretical cycle P _i and on its efficiency is exerted by the compression ratio ε, therefore, Figures 9 and 11 show in detail the effect

on P _i and η _i for various ε = 8 ÷ 14.

As a result of theoretical and theoretical studies, it was found that for theoretical cycles with isochoric supply and removal of heat, there is an optimal amount of water, which, when supplied to the working fluid, at the beginning of compression, the greatest efficiency η _i and average pressure P _{i of} theoretical cycles are obtained. The optimal amount of water depends on the degree of compression. For large degrees of compression, it is closer to 0.375, and for lower degrees closer to 0.5. A smaller amount of water is characteristic of higher T _a and low P _a , and a larger amount must be supplied at lower T _a and high P _a .

при коэффициентах избытка воздуха α=0,93; 1,0; 1,08 [Звонов В.А. Токсичность двигателей внутреннего сгорания. - 2-е изд., перераб. - М.: Машиностроение, 1981. - 160 с.], а также рассчитывались параметры и технико-экономические показатели двигателя МЗМА-407 при различных количествах добавляемой воды во впускной воздух

по математической модели. Результаты расчетов приведены на графиках (фиг.12 и 13).To experimentally confirm the theoretical results, we used experimental data on the change in the specific effective fuel consumption g _e , the average effective pressure P _e depending on the amount of sprayed water in the combustible mixture

with coefficients of excess air α = 0.93; 1.0; 1,08 [Zvonov V.A. Toxicity of internal combustion engines. - 2nd ed., Revised. - M.: Mechanical Engineering, 1981. - 160 p.], And also calculated the parameters and technical and economic indicators of the MZMA-407 engine with various amounts of added water into the intake air

according to the mathematical model. The calculation results are shown in the graphs (Fig.12 and 13).

в горючую смесь коэффициент наполнения η_V сначала увеличивается, а затем уменьшается (фиг.12). Повышение η_V происходит благодаря уменьшению температуры горючей смеси в конце впуска Т_а, несмотря на увеличение коэффициента остаточных газов γ, до

. Затем этот коэффициент начинает убывать из-за более быстрого увеличения γ при небольшом повышении Т_a. Повышение коэффициента остаточных газов γ связано с уменьшением температуры остаточных газов Т_r при всех значениях α. Для α=0,93 коэффициенты остаточных газов γ (кривая 1) получаются больше, чем для α=1 (кривая 2), из-за меньших температур остаточных газов Т_r, несмотря на более высокие коэффициенты наполнения η_V. При α=1,08 кривая 3 остаточных газов располагается выше, чем при α=1 (кривая 2) и α=0,93 (кривая 1), из-за более низких значений Т_r, чем у кривой 2 (α=1), и более меньших значений коэффициентов наполнения η_V (кривая 3), чем у кривых 1 (α=0,93) и 2 (α=1). Уменьшение температуры остаточных газов с увеличением

связано с уменьшением максимальной температуры газов T_Z и температуры газов в конце расширения T_b.With an increase in the amount of water injected

into the combustible mixture, the filling coefficient η _V first increases and then decreases (Fig. 12). An increase in η _V occurs due to a decrease in the temperature of the combustible mixture at the end of the inlet T _a , despite the increase in the coefficient of residual gases γ, to

. Then this coefficient begins to decrease due to a faster increase in γ with a slight increase in T _a . An increase in the coefficient of residual gases γ is associated with a decrease in the temperature of the residual gases T _r for all values of α. For α = 0.93, the residual gas coefficients γ (curve 1) are obtained more than for α = 1 (curve 2), due to lower temperatures of the residual gases T _r , despite the higher filling factors η _V. For α = 1.08, curve 3 of the residual gases is higher than for α = 1 (curve 2) and α = 0.93 (curve 1), due to lower T _r values than curve 2 (α = 1 ), and lower values of the filling coefficients η _V (curve 3) than curves 1 (α = 0.93) and 2 (α = 1). Decrease in temperature of residual gases with increase

due to a decrease in the maximum gas temperature T _Z and gas temperature at the end of expansion T _b .

до 0,5÷0,75 уменьшаются, а величины индикаторного η_i и эффективного КПД повышаются при всех α=0,93; 1,0; 1,08 (фиг.13). При

g_e начинает возрастать, a η_i и η_е убывать. Характер изменений кривых q_e, η_I,

, в основном, определяется значениями максимальных температур T_Z и давлений P_Z газов, давлениями горючей смеси в конце сжатия Р_C, коэффициентами использования теплоты υ_Z в в.м.т., а также величинами показателей политроп сжатия n_C и расширения n_P. Несмотря на незначительное влияние процесса сгорания на продувку и наполнение цилиндров характер изменения кривых η_i и

получился очень даже похожим на кривые изменения теоретического КПД термодинамического цикла с изохорными подводом и отводом теплоты. Разница заключается лишь в оптимальных значениях этих КПД, которые для термодинамического цикла получились в диапазоне

в зависимости от ε, Р_а, Т_а, а для действительного цикла в области

в зависимости от α. При таком пологом изменении этих кривых такое расхождение может быть как за счет точности проведения экспериментов, так и расчетов, тем более, что у них имеется общая цифра

при одинаковых степенях сжатия ε=8 и коэффициентах избытка воздуха α=1. Приращения абсолютных значений теоретического и индикаторного КПД в зависимости от

у этих циклов получаются также почти одинаковыми. Разница, конечно, есть, но и она не принципиальна и на порядок меньше погрешностей экспериментального определения η_i. Абсолютные значения КПД теоретического цикла получились значительно больше индикаторного КПД реального цикла из-за отсутствия тепловых потерь в систему охлаждения, химической неполноты сгорания топлива, а также сгоранию топлива в в.м.т. в этом цикле.The value of the specific effective fuel consumption g _e with increasing

up to 0.5 ÷ 0.75 decrease, and the values of indicator η _i and effective efficiency increase for all α = 0.93; 1.0; 1.08 (Fig.13). At

g _e begins to increase, a η _i and η _e decrease. The nature of the changes in the curves q _e , η _I ,

mainly determined by the values of the maximum temperatures T _Z and pressures P _{Z of the} gases, the pressures of the combustible mixture at the end of compression Р _C , the heat utilization factors υ _Z in bhp, as well as the values of the polytropic parameters of compression n _C and expansion n _P . Despite the insignificant effect of the combustion process on the purge and filling of the cylinders, the nature of the change in the curves η _i and

it turned out to be very similar to the curves of changes in the theoretical efficiency of the thermodynamic cycle with isochoric supply and removal of heat. The difference lies only in the optimal values of these efficiency, which for the thermodynamic cycle were in the range

depending on ε, P _a , T _a , and for the actual cycle in the field

depending on α. With such a gentle change in these curves, such a discrepancy can be due to the accuracy of the experiments and calculations, especially since they have a common figure

at the same compression ratios ε = 8 and air excess coefficients α = 1. Increments in the absolute values of theoretical and indicator efficiency depending on

these cycles are also almost identical. There is, of course, a difference, but it is not fundamental and is an order of magnitude smaller than the errors of the experimental determination of η _i . The absolute values of the efficiency of the theoretical cycle turned out to be much more than the indicator efficiency of the real cycle due to the absence of heat losses in the cooling system, chemical incompleteness of fuel combustion, as well as fuel combustion in bhp. in this cycle.

The results of the studies allow us to increase the efficiency of real cycles, not only due to an increase in the amount of work performed as a result of adding water to the combustible mixture and lowering the maximum combustion temperature, but also due to an increase in the compression ratio, without increasing the octane number of the fuel and increasing the excess air coefficient to α = 1.1 ÷ 1.2, exceptions from the designs of engines of the cooling system, and also due to the utilization of the heat of the spent oil and combustion products.

It was discussed above how much water needs to be sprayed in the intake air so that the effective efficiency of the gasoline engine is greatest when the water droplets are heated by the end of compression to the boiling temperature T _water.k . This temperature is achieved by changing the spray quality, which is characterized by the average droplet size (Internal combustion engines: Theory of piston and combined engines. A textbook for universities on a specialty "Internal combustion engines" / D.N. Vyrubov, N.A. Ivaschenko, V.I. Ivin and others. - M.: Mechanical Engineering, 1983.- 372 p., Ill.). Here under the average droplet size d _k.av. characterizing the quality of spraying water, we understand the diameter of the droplets, when calculated according to which they are heated by the end of compression to T _water.k , while the pressure P _s and temperature T _{with the} combustible mixture at this time become the smallest, and the cycle work is the largest for a given or optimal amount of sprayed water. To determine this droplet size, the amount of heat was determined, which is transferred to water droplets with dimensions d _{K. CP} heat transfer from the working fluid during the time dt

where: F _K = πd ² _{K. CP} - surface of a droplet of size d _{K. CP} , m ² ;

i _K is the number of drops of size d _{K. CP} ;

- средний коэффициент теплоотдачи за

;

- average heat transfer coefficient per

;

- температура сжатия рабочего тела в зависимости от времени t, К;

- the compression temperature of the working fluid, depending on the time t, K;

- температура капли воды в зависимости от времени t, К.

- temperature of a drop of water depending on time t, K.

As well as the amount of heat that is perceived by drops of water when they are heated at AT _K

;where: _{WATER. SR} - average heat capacity of water droplets with size d _{K. SR} when their temperature changes by dТ _K ,

;

- масса капли размером d_К.СР, кг;

- the mass of a drop of size d _{K. CP} , kg;

ρ _{WATER. SR} is the average density of water droplets of size d _{K. SR} when their temperature changes by dT _K , kg / m ³ .

Integrating equations 36 and 37, transforming and equating them, we obtain the expression for determining d _{K. CP}

- среднее значение критерия Нуссельта за Δφ=180° п.к.в.;Where:

- the average value of the Nusselt criterion for Δφ = 180 ° p.c.

T _VOD.N - initial temperature of the sprayed water in the air, K;

ω is the crankshaft rotation frequency, 1 / s.

и давления

при сжатии рабочего тела предложены уравнения (39) и (40)To determine the current temperature

and pressure

when compressing the working fluid, equations (39) and (40) are proposed

от количества поданного топлива, при различных коэффициентах избытка воздуха α и кондуктивном теплообмене этих капель с горючей смесью (N_n=2). Результаты расчетов сведены в таблицу. В эту таблицу занесены еще значения давления Рс и температуры Тc горючей смеси в конце сжатия, а также среднего размера капель d_К.СР двигателя МЗМА-407, работающего при ω=2600 об/мин, различных

и α, давление Рс и температура Тс горючей смеси в конце сжатия.According to the developed method, using numerical methods, d _K.СР were calculated for the MZMA-407 engine at n = 2600 rpm depending on the proportion of injected water

the amount of fuel supplied, at various coefficients of excess air α and the conductive heat transfer of these droplets with a combustible mixture (N _n = 2). The calculation results are summarized in the table. This table also _{contains the} values of the pressure Pc and temperature Tc of the combustible mixture at the end of compression, as well as the average droplet size d _{K. CP of} the MZMA-407 engine operating at ω = 2600 rpm, various

and α, pressure Pc and temperature Tc of the combustible mixture at the end of compression.

и αTable
The values of the pressure Pc and temperature Tc of the combustible mixture at the end of compression, as well as the average droplet size d _{K. CP of} the MZMA-407 engine operating at ω = 2600 rpm, various

and α

α 0.93 1,0 1,08 one 2 3 four 5 0.25 d _{K. CP} , μm 233 256 272 Pc · 10 ⁵ , Pa 11.2 10.8 10.3 Ts, K 570 590 600 0.50 d _{K. CP} , μm 195 220 238 Pc · 10 ^-5 , Pa 10.3 10.0 9.65 Ts, K 525 540 550 0.75 d _{K. CP} , μm 160 190 212 Pc · 10 ^-5 , Pa 9.7 9.2 9.0 Ts, K 480 510 520 1,0 d _{K. CP} , μm 133 163 186 Pc · 10 ^-5 , Pa 9.0 8.8 8.6 Ts, K 460 475 490 1,08 d _{K. CP} , μm 108 140 165 Pc · 10 ^-5 , Pa 8.5 8.3 8.2 Ts, K 435 455 480

, тем меньше Рс и Тс, а также требуемый размер капель d_К.СР из-за уменьшения Рс и Тс. Чем меньше Рс при Тс=const, тем на меньшую температуру должны быть нагреты однородные капли, что возможно при меньшей их поверхности, а следовательно, большем d_К.СР. С уменьшением Тс при Рc=const тепловой поток к каплям уменьшается, его уменьшение компенсируется увеличением поверхностей этих капель, а следовательно, уменьшением d_К.СР. Следовательно, на уменьшение d_К.СР влияет не только уменьшение Рс с увеличением

, но и уменьшение Тс. С увеличением коэффициента избытка воздуха Рс уменьшается, а Т_с увеличивается, а требуемый размер капель d_К.СР растет из-за преобладающего влияния увеличения Тс. При α=1 и наивыгоднейшем

размер капель d_К.СР получается 190÷220 мкм. Для того чтобы получить такой размер капель в горючей смеси достаточно вытекающую струю воды из распылителя распыливать потоком воздуха в диффузоре [Автомобильные двигатели / В.М.Архангельский, М.М.Вихерт, А.Н.Воинов и др. - М.: Машиностроение, 1967. - 496 с.: ил.].The table shows that the more

, the smaller Pc and Tc, as well as the required droplet size d _{K. CP} due to a decrease in Pc and Tc. The smaller Pc at Tc = const, the homogeneous droplets should be heated to a lower temperature, which is possible with a smaller surface and, consequently, a larger d _{K. CP} . With a decrease in Tc at Pc = const, the heat flux to the drops decreases, its decrease is compensated by an increase in the surfaces of these drops, and therefore, a decrease in d _{K. CP} . Therefore, a decrease in d _{K. CP is} affected not only by a decrease in Pc with increasing

, but also a decrease in Tc. With an increase in the coefficient of excess air, Pc decreases and T _c increases, and the required droplet size d _{K. CP} increases due to the predominant influence of an increase in Tc. For α = 1 and the best

the droplet size d _{K. CP} is 190 ÷ 220 microns. In order to obtain such a droplet size in a combustible mixture, a sufficiently flowing stream of water from a spray gun should be sprayed with a stream of air in a diffuser [Automotive engines / V.M. Arkhangelsky, M.M. Vichert, A.N. Voinov et al. - M.: Mechanical Engineering , 1967. - 496 p.: Ill.].

Thus, already at the stage of designing a new engine, d _K.СР can be calculated for it, and the diameter of the nozzle opening of the atomizer can be determined from it (Fuel supply and atomization in diesel engines. Astakhov IV, Trusov VI, Khachiyan A .S. Et al. - M.: Mechanical Engineering, 1971. - 359 p.: Ill.). Then, this diameter is specified when testing on an industrial or prototype at the lowest pressure and temperature at the end of compression and the highest indicator and effective efficiency.

The size of the atomizer can be determined even empirically on a test or industrial model by injecting water through various atomizers into the intake air and scrolling the piston engine by the electric motor at the lowest pressure and temperature of the combustible mixture at the end of compression. On a running engine, it can also be installed only by the lowest pressure and temperature of the working fluid at the beginning of combustion and the highest indicator or effective efficiency. In this case, the average d _{K. CP} can also be established by measuring the diameter of homogeneous droplets in the torches formed by spraying water through sprayers of various sizes. If inhomogeneous droplets form in the flares, then d _{K. CP} can be determined by calculation according to equation (38) and the known experimental minimum pressure and temperature at the end of compression.

необходимо добавлять во впускной воздух, чтобы не возникала в них детонация. Известно, что основными факторами, оказывающими наиболее значительное влияние на возникновение детонации в бензиновых двигателях являются максимальные температура T_Z и давление P_Z сгорания. Были проведены исследования с помощью разработанной авторами математической модели и экспериментальных данных, какое должно быть снижение Т_Z при повышении P_Z. При этом анализировались экспериментальные данные различных исследователей, в которых изменялись количество подаваемой воды во впускной воздух, давление и температура надувочного воздуха, степень сжатия, коэффициент избытка воздуха, угол опережения зажигания и нагрузка (Ефремов П.К. К вопросу о дополнительном питании тепловых двигателей водой. - В кн.: Защита воздушного бассейна от загрязнения токсичными выбросами транспортных средств. - Т.1. - Харьков, 1977. - С.221-261, а также Добрынин А.А. Детонация в двигателях, 1949. - 58 с.: ил.). Проведенные исследования позволили построить графики понижения максимальной температуры газов на единицу повышения максимального давления сгорания

в зависимости от значения T_Z (фиг.14), потерь теплоты в систему охлаждения q_ОХЛ (фиг.15) и с выпускными газами q_ВЫП (фиг.16). С увеличением значения T_Z средние значения

уменьшаются (фиг.14). Большие значения

относятся к экспериментам с меньшими значениями q_ОХЛ и с большими величинами q_ВЫП, т.е., для опытов с повышенной теплонапряженностью двигателя. Меньшие значения

характерны для экспериментов, в которых значения q_ОХЛ получались наибольшими, а значения q_ВЫП наименьшими, т.е. для тех случаев, в которых теплонапряженность двигателя была наименьшей. С уменьшением значений q_ОХЛ, то есть с возрастанием теплонапряженности двигателя, значения

увеличиваются (фиг.15), причем большие значения относятся к опытам, в которых значения T_Z были наименьшими, меньшие величины к экспериментам с более высокими T_Z. С возрастанием значений q_ВЫП значения

возрастают (фиг.16) из-за повышения теплонапряженности двигателя, причем большие значения относятся к экспериментам, в которых значения T_Z были наименьшими, а меньшие величины к опытам с большими значениями T_Z.To realize a high compression ratio in combined gasoline engines, you need to know how much water

must be added to the intake air so that knocking does not occur in them. It is known that the main factors that have the most significant effect on the occurrence of detonation in gasoline engines are the maximum temperature T _Z and pressure P _{Z of} combustion. Studies were conducted using the mathematical model developed by the authors and experimental data, which should be a decrease in T _Z with increasing P _Z. In this case, the experimental data of various researchers were analyzed, in which the amount of water supplied to the inlet air, the pressure and temperature of the charge air, the compression ratio, the excess air coefficient, the ignition timing and the load (Efremov P.K., On the issue of additional supply of heat engines to water) - In the book: Protection of the air basin from pollution by toxic emissions of vehicles - T.1. - Kharkov, 1977. - P.221-261, as well as Dobrynin A.A. Detonation in engines, 1949. - 58 p. : ill.). The studies performed allowed us to construct graphs for lowering the maximum gas temperature per unit of increase in the maximum combustion pressure

depending on the value of T _Z (Fig. 14), heat loss to the cooling system q _OHL (Fig. 15) and with the exhaust gases q _RH (Fig. 16). With increasing values of T _Z average values

decrease (Fig. 14). Big values

relate to experiments with lower values of q _OHL and with large q values of _RPC , i.e., for experiments with increased thermal stress of the engine. Lower values

typical for experiments in which the q _OHL values were obtained by the largest, and the _qQR values were the smallest, i.e. for those cases in which the thermal stress of the engine was the smallest. With a decrease in q _OHL values, i.e., with an increase in the thermal stress of the engine, the values

increase (Fig. 15), with larger values referring to experiments in which the values of T _Z were the smallest, lower values to experiments with higher T _Z. With increasing values of q values _SAT

increase (Fig. 16) due to an increase in the thermal stress of the engine, with larger values referring to experiments in which the values of T _Z were the smallest and smaller values to experiments with large values of T _Z.

The obtained graphs in Figs. 14, 15 and 16 will allow further calculations to be made to increase the compression ratio of gasoline engines when they are created with a two-phase working fluid and to exclude the cooling system from them without increasing the octane number of the fuel, as well as increasing the heat stress due to the injection of water into the combustible mixture.

In order to calculate the parameters and indicators of a combined gasoline engine with converted (s) gas (s) cylinder (s) into steam (s), the authors developed a mathematical model for the thermal calculation of the cycle of gasoline engines, supplemented by equations to determine the amount of steam produced and calculate the work, performed by him in the steam cylinder (s).

The amount of steam generated in the steam generator m _pairs can be determined in kg / s from the heat balance equation

where: T _PA - the temperature of the generated steam, K;

C _PAR - the heat capacity of steam at constant pressure and T _PAR , J / (kg · K);

T _WATER is the temperature of the water entering the steam generator, K;

With _WATER is the heat capacity of the water entering the steam generator, J / (kg · K);

G _m - second fuel consumption, kg / s;

α is the coefficient of excess air;

L ₀ - theoretically necessary amount of air for combustion of 1 kg of fuel, kg / kg;

- относительное количество воды, добавляемой в горючую смесь, кг/кг;

- the relative amount of water added to the combustible mixture, kg / kg;

T _p - temperature of the combustion products entering the steam generator, K;

C _PR.C - heat capacity of the combustion products entering the steam generator at constant pressure and T _p , J / (kg · K);

T _UH.G - temperature of the combustion products behind the steam generator, K;

C _UH.G - the heat capacity of the exhaust products of combustion at constant pressure and T _UH.G , J / (kg · K);

m _WATER is the amount of water supplied to the steam generator, kg / s.

If in equation (41), instead of m _WATER, substitute m _PAIR , then from this equation we can determine

To determine the vapor pressure in the steam (s) cylinder (s) at the beginning of expansion, we write the equation of material balance

where: V _N.P - the volume of steam in the cylinder (s) at the beginning of expansion, after closing the inlet steam valve φ ₁ , m ³ ;

ρ _{N. R} - vapor density in the cylinder (s) at the beginning of expansion, kg / m ³ ;

i _{C. n} is the number of cylinders converted to work on a pair;

ω is the crankshaft rotation frequency, 1 / s.

We substitute the value of vapor density in equation (43)

, тогда получим

then we get

where: R _PAR - gas constant for steam, J / (kg · K).

The specific volume of steam at the end of the expansion is determined

where: υ _N.Р = 1 / ρ _N.Р - specific volume of steam at the beginning of expansion, m ³ / kg;

υ _K.R - specific volume of steam in the cylinder (s) at the end of expansion at the time of opening of the steam exhaust valve φ ₂ , m ³ .

According to the specific volume of steam at the end of the expansion and its adiabatic expansion in the cylinder, we find the temperature T _K.P and the pressure P _K.P at the end of the expansion according to the tables for water vapor.

The indicator work performed by the steam in the cylinder before the opening of the exhaust steam valve is determined by the equation

where: h _K.R and h _N.R - vapor enthalpy at the beginning and at the end of the expansion process in the cylinder, J / (kg · K);

R _N.R and R _K.R — vapor pressure at the beginning and at the end of the expansion process, Pa;

υ _N.Р and υ _К.Р - specific volumes of steam at the beginning and at the end of expansion, kg / m ³ .

The values of h _K.P and h _N.P are found in the tables for water vapor.

To determine the work performed by the steam in the steam cylinder during the pre-release period of L _O.V. (from φ ₂ to 180 ° p.c.), as well as the work spent by the piston on expelling the steam when it moves from the engine to m.t. (from 180 to 360 gr.p.k.v), we write the equations of state and material balance

и

- начальные значения давления и температуры пара в конце расширения в Па и К;Where:

and

- initial values of pressure and temperature of steam at the end of expansion in Pa and K;

- начальное значение количества пара в цилиндре в конце расширения в кг, принимается равным m_ПАР;

- the initial value of the amount of steam in the cylinder at the end of the expansion in kg, is taken equal to m _PAIR ;

where: dt is the change in time during which steam expires from the cylinder

,

,

where: Δφ - change in the angle of rotation of the crankshaft, gr. p.c.

- средняя плотность пара, вытекающего из цилиндра, на участке Δφ, кг/м³;

- the average density of the vapor flowing from the cylinder in the area Δφ, kg / m ³ ;

- среднее проходное сечение выпускного клапана на участке Δφ, м²;

- the average flow cross section of the exhaust valve in the area Δφ, m ² ;

- средняя скорость движения пара в сечении выпускного клапана на участке Δφ, м/с.

- the average speed of steam in the cross section of the exhaust valve in the area Δφ, m / s.

рассчитывается по средней плотности пара в этом сечении, разности средних давлений пара в цилиндре и в конденсаторе, а также текущему сопротивлению этого клапана.The average speed of steam in the cross section of the exhaust valve

calculated by the average vapor density in this section, the difference between the average vapor pressures in the cylinder and in the condenser, as well as the current resistance of this valve.

определяется по уравнению состояния и известному

, которое находится по начальному

и текущему

удельным объемам пара и таблицам для водяного пара при адиабатном расширении.Current cylinder pressure

determined by the equation of state and known

which is at the initial

and current

specific volumes of steam and tables for water vapor with adiabatic expansion.

The work performed by the steam in the cylinder from the moment the exhaust valve is opened (φ ₂ ) until the piston reaches m.t. (180 gr.p.k.)

where: P _SR.n is the average vapor pressure in the cylinder in the n section, m ³ ;

ΔV _n is the change in the volume of the cylinder in the n section, m ³ .

The work spent by the engine on pushing steam out of the cylinder as it moves from the engine (180 gr.p.k.) to the upper (360 gr.p.k.)

The work spent by the engine on overcoming the friction forces of the piston in the steam cylinder,

where: P _MEX - average pressure of mechanical losses;

V _{h is the} working volume of the cylinder, m ³ .

The total indicator work performed by the steam in one of the cylinders,

The efficient work done by steam in one of the cylinders,

where: L _P.N - the work spent by the engine on the drive of the feed pump.

The value of L _{N. H is} determined by the formula

where: P _NAS - water supply pressure equal to P _PAR , Pa;

η _NAS - pump efficiency;

ω _NAS - the frequency of rotation of the pump shaft, equal to the rotational speed of the crankshaft ω, 1 / s.

The effective power of the combined engine without turbo expander

Specific effective fuel consumption of a combined engine without a turboexpander

Effective efficiency of a combined engine without a turbo-feeder

In the presence of a turboexpander, the back pressure in the exhaust manifold P _P will change and the pressure P _K and the charge air temperature T _K will appear to calculate the real cycle of the gasoline engine.

Compressor charge air pressure P _{K is} determined by the equation

where: P _O - air pressure at the inlet to the compressor, Pa;

K _K is the adiabatic index of air compression in the compressor;

R _WHO - gas constant for air, J / (kg · K);

T _About - the air temperature at the inlet to the compressor, K;

G _K - air flow through the compressor, kg / s;

N _T - net power of the turbine, W;

η _K - adiabatic efficiency of the compressor.

Charge-air temperature T _K compressor Calc'd

The air flow through the compressor will be

where: G _T - fuel consumption per second, kg / s.

Net power of the turbine

where: η _T is the effective efficiency of the turbine;

- степень понижения давления газов в турбине;

- the degree of decrease in gas pressure in the turbine;

P _R - gas pressure in front of the turbine, Pa;

P _R.O - gas pressure at the outlet of the turbine, Pa;

T _P is the temperature of the gases in front of the gas turbine, K;

G _T.PR.G.G = G _PR.G. G _m - gas flow through the turbine, kg / s.

The gas temperature behind the gas turbine

where: η _T.AD - adiabatic efficiency of the turbine.

уменьшается. Происходит это из-за того, что увеличивается температура горючей смеси в конце впуска и в конце сжатия, а коэффициент наполнения уменьшается. Значительное уменьшение эффективной мощности происходит из-за уменьшения количества сжигаемого топлива в цилиндре при неизменном коэффициенте избытка воздуха.To confirm the industrial applicability and efficiency of the combined engine with a two-phase working fluid, the parameters and parameters of the cycles of this engine were calculated at a normal outdoor temperature t _WHO = 20 ° C and its temperature t _WHO = 40 ° C without a turbine expander in the presence of a throttle valve at the outlet and with a turboexpander (Fig.17, 18, 19). The modern gasoline engine of the Zavolzhsky engine plant ZMZ-4062.10 was chosen as the base engine. The calculations of this engine were performed at t _WHO = 40 ° C, and its parameters and indicators are presented in Figs. 17, 18, 19. At this temperature, at all operating modes, its effective power is much lower than at t _WHO = 20 ° C, and at n = 5000 rpm, this decrease is 8 kW (7.7%). Its effective efficiency also decreases throughout the entire external speed characteristic, and this decrease at n = 5000 rpm is 0.5%. In this case, the maximum combustion temperature T _Z increases, and the pressure

decreases. This is due to the fact that the temperature of the combustible mixture at the end of the inlet and at the end of compression increases, and the filling coefficient decreases. A significant decrease in effective power occurs due to a decrease in the amount of fuel burned in the cylinder with a constant coefficient of excess air.

The engine is equipped with a steam generator, converting one of the cylinders to work on steam, increasing its compression ratio to ε = 14, increasing the coefficient of excess air from α = 0.9 to α = 1.1, suppressing the resulting detonation in the engine by supplying water to the combustible mixture, as well as its operation without a throttle valve at the outlet and an outdoor temperature of 20 ° С, its effective efficiency significantly increases up to 0.592 at ω = 1000 rpm and up to 0.506 at ω = 5000 rpm, and effective power up to 127 kW at ω = 5000 rev / min, i.e. there is an increase in η _e 1.85 ÷ 1.87 times, and efficient th power 1,22 ÷ 1,23 times across the outside speed. Minimal increase in N _e N _e in comparison occurs due to power conversion and one cylinder gasoline engine to run on a pair.

These results were obtained despite the fact that during the calculations, the resistance of the exhaust tract due to the presence of a steam generator increased by 1.2 times, and the value of the heat utilization factor in bm decreased due to an increase in the coefficient of excess air from 0.9 to 1.1 by 1.14 ÷ 1.18 times, according to studies of traditional and modern engines of the ZMZ-406.10 family.

по сравнению с

и T_Z базового двигателя составило при ω=5000 об/мин

, ω=4000 об/мин

, а при ω=3000 об/мин

, что допустимо для бездетонационной работы двигателя. Благодаря увеличению степени сжатия и коэффициента избытка воздуха произошло у него повышение среднего индикаторного давления Р_i во всем диапазоне внешней скоростной характеристики (кривая 2 на фиг.18) по отношению к P_i=f(ω) базового двигателя как при t_ВОЗ=20°С, так и t_ВОЗ=40°С (кривая 1 на фиг.18). Кривая 2 Р_Н.Х=f(ω) комбинированного двигателя (фиг.18)) получилась почти одинаковой с кривой 1 Р_Н.Х=f(ω) базового двигателя как при t_ВОЗ=20°С, так и t_ВОЗ=40°С (фиг.18).Due to the addition of water to the combustible mixture of the combined engine and an increase in its excess air coefficient from α = 0.9 to α = 1.1, there was a significant decrease in T _{Z of} this engine over the entire external speed characteristic, despite an increase in the compression ratio from ε = 9.3 to ε = 14 (curve 2 in FIG. 17). The maximum combustion pressure increased (curve 2 in FIG. 18) due to an increase in compression end pressure due to increases in compression ratio and fill factor, despite a decrease in T _Z. Decrease in maximum combustion temperature per unit of increase in maximum combustion pressure

compared with

and T _{Z of the} base engine was at ω = 5000 rpm

, ω = 4000 rpm

, and at ω = 3000 rpm

, which is acceptable for detonation-free engine operation. Due to the increase in the compression ratio and the coefficient of excess air, he had an increase in the average indicator pressure P _i in the entire range of the external speed characteristics (curve 2 in Fig. 18) with respect to P _i = f (ω) of the base engine as at t _WHO = 20 ° C and t _WHO = 40 ° C (curve 1 in Fig. 18). Curve 2 P _N.X = f (ω) of the combined engine (Fig. 18)) turned out to be almost the same as curve 1 P _N.X = f (ω) of the base engine both at t _WHO = 20 ° C and t _WHO = 40 ° C (Fig. 18).

у этого двигателя как при t_ВОЗ=20°С, так и t_ВОЗ=40°С (кривые 2 и 3 на фиг.19) кривая T_Z=f(ω) при t_ВОЗ=40°С (на фиг.17) расположилась выше аналогичной кривой, рассчитанной при t_ВОЗ=20°С, кривая

, наоборот, ниже по сравнению с кривой

при t_ВОЗ=20°С. Такие изменения T_Z и P_Z связаны с уменьшением коэффициента наполнения η_V и увеличением температуры конца впуска Т_а при повышении температуры наружного воздуха. При этом снижение максимальной температуры сгорания на единицу повышения максимального давления сгорания составило по сравнению с

и T_Z базового двигателя при ω=5000 об/мин

, ω=4000 об/мин

и ω=3000 об/мин

. Снижение температуры ΔТ_Z произошло на значительно большую величину, чем требуется для работы двигателя без детонации. Такие высокие

получились, в основном, из-за низких

в результате уменьшения коэффициента наполнения.At an outside air temperature t _WHO ≤20 ° С, the problem of condensate cooling with this air can be solved without a significant increase in the heat transfer surface, since the temperature difference between them will be at least 40 ° С due to condensation of water vapor at a temperature of more than 60 ° С in the absence of a turbine expander and throttle on release. With increasing t _WHO > 20 ° C, the amount of transferred heat from condensate to air begins to decrease due to a decrease in the temperature difference between them. To maintain this temperature difference between these coolants, it is necessary to increase the condensation temperature of water vapor in the condenser t _P by increasing the pressure P _{P in it} , introducing a turbo expander or throttle valve into the design of the exhaust system. With an increase in t _WHO ≥40 ° C, it is necessary to increase t _P ≥80 ° C by increasing the pressure P _P ≥2.5 · 10 ⁵ Pa. Such an increase in pressure, first of all, affects the course of the processes of the actual cycles and the increase in the operation of pumping strokes, and consequently, the deterioration of the technical and economic indicators of the combined engine. Therefore, the calculations were made of the working processes of this engine at P _P = 2.5 · 10 ⁵ Pa and t _WHO = 40 ° C with a throttle valve at the outlet and with a turboexpander. On Fig. Curve 3 shows the change in the effective power of the combined engine with a throttle at the outlet according to the external high-speed characteristic. This curve turned out to be lower than the curve N _e = f (ω) of the ZMZ-4062 base engine (curve 1 of Fig. 17), calculated at t _WHO = 40 ° C. A decrease in the fill factor due to an increase in the coefficient of residual gases and an increase in the temperature of the combustible mixture at the end of the inlet, as well as a significant increase in the operation of the pump strokes (curve 3 of Fig. 18), affected the decrease in power of this engine. The decrease in N _e at a crankshaft speed of 5000 rpm occurred by 6.5 kW compared to the N _{e of the} base engine at t _WHO = 40 ° C. This power reduction can be eliminated if necessary by a short-term decrease in P _P. Despite the decrease in the effective power of the combined engine at t _WHO = 40 ° C and the presence of a throttle valve at the outlet, its effective efficiency significantly increased throughout the entire external speed characteristic (curve 3 of Fig. 17) relative to η _е = f (ω) of the base engine as at t _WHO = 20 ° C, and t _WHO = 40 ° C. With the same change

for this engine, both at t _WHO = 20 ° C and t _WHO = 40 ° C (curves 2 and 3 in Fig. 19), the curve T _Z = f (ω) at t _WHO = 40 ° C (in Fig. 17 ) is located above a similar curve calculated at t _WHO = 20 ° C, the curve

conversely, lower than the curve

at t _WHO = 20 ° C. Such changes in T _Z and P _Z are associated with a decrease in the fill factor η _V and an increase in the temperature of the inlet end T _a with increasing outdoor temperature. In this case, a decrease in the maximum combustion temperature per unit of increase in the maximum combustion pressure amounted to

and T _{Z of the} base engine at ω = 5000 rpm

, ω = 4000 rpm

and ω = 3000 rpm

. The temperature ΔТ _Z decreased by a significantly larger amount than is required for engine operation without detonation. So tall

turned out mainly due to low

as a result of a decrease in fill factor.

As calculations showed, the problem of water regeneration from combustion products at t _WHO ≥40 ° C, even without a significant complication of the design, is solved by increasing the effective efficiency at ω = 5000 rpm by 1.55 times, ω = 4000 rpm at 1 51 times and ω = 3000 rpm 1.45 times with respect to η _{e of the} base engine.

The use of a turboexpander instead of a throttle valve for the regeneration of water from combustion products at high outdoor temperatures not only reduces the effective power of the combined engine, but also increases it (curve 4 of Fig. 17). The increase in effective power at ω = 5000 rpm was 49.2 kW or 51.1% with respect to the N _{e of the} base engine at t _WHO = 40 ° C.

Its effective efficiency also increases significantly (curve 4 of Fig. 17), but it turns out to be less than that of a combined engine operating at t _WHO = 20 ° C without increasing the pressure in the condenser. An increase in effective efficiency compared to η _{e of the} base engine is obtained at ω = 5000 rpm by 1.73 times, and at ω = 4000 rpm and ω = 3000 rpm by 1.75 times.

(кривая 4 фиг.19). Дальнейшее снижение T_Z на этих режимах повлияет на снижение коэффициента использования теплоты в в.м.т. υ_Z. При ω=1000 об/мин эта температура еще больше снижается (кривая 4 на фиг.17) из-за уменьшения υ_Z в результате падения ω несмотря на уменьшение

до 0,55.The maximum combustion temperature with this equipment in the range ω = 2000-5000 rpm is extremely low, about 2200 K (curve 4 of Fig. 17) when water is supplied to the combustible mixture

(curve 4 of Fig. 19). A further decrease in T _Z in these modes will affect a decrease in the heat utilization coefficient in bhp. υ _Z. At ω = 1000 rpm, this temperature decreases even more (curve 4 in Fig. 17) due to a decrease in υ _Z as a result of a drop in ω despite a decrease

up to 0.55.

The maximum combustion pressure of the engine supercharging turbo expander becomes extremely high in the whole range ω, reaching at ω = 4000 rpm / min P _Z = 111,7 · 10 ⁵ Pa (18). This increase in P _Z is mainly due to changes in P _K (curve 4 in FIG. 19) and the fill factor η _V as a function of ω.

The pressure and temperature of the charge air were determined from the given η _T and η _K depending on ω (curve 4 in Fig. 19) at constant P _P = 2.5 · 10 ^-5 Pa and t _P = 80 ° C.

гр/Па, которые расположились близко к нижней границе допустимых значений этого показателя.The value of lowering the maximum combustion temperature per unit of increase in the maximum combustion pressure was obtained in the range of change ω = 2000 ÷ 5000 rpm

g / Pa, which are located close to the lower boundary of the permissible values of this indicator.

The average pressure of the pumping strokes of the combined engine with a turboexpander (curve 4 in Fig. 18) turned out to be 1.25–1.3 times lower than with a throttle valve (curve 3) due to engine boost. Curve 4 of the average indicator pressure passed above all curves 1, 2 and 3 (Fig. 18), mainly due to the supply of air to the cylinders under pressure.

According to the combined engine with a turboexpander, the following conclusion can be drawn that, at high outdoor temperatures, it allows water to be released from the combustion products with a lower decrease in effective efficiency than when using a throttle valve. At the same time, its effective power still increases significantly in the entire range of external speed characteristics. When deforming the power of this engine by reducing the rotational speed of the collector shaft to the level of effective power of the combined engine operating at t _WHO = 20 ° C (curve 2 in Fig. 17), their effective efficiency is almost compared, with the same N _e . With even more de-stressing of its power to the level of effective power of the base engine operating at an outdoor temperature of 40 ° C, its effective efficiency in nominal mode can increase to 0.54.

Diesel engines differ significantly from gasoline engines mainly in that heterogeneous combustion and self-ignition of the fuel occurs in these engines. They have a higher compression ratio compared with the compression ratios of gasoline engines, and therefore, higher efficiency. The combustion process takes place with a higher coefficient of excess air in the nominal operating mode, and with a decrease in load, this coefficient increases even more, which is associated with high-quality power control. At high excess air ratios, lower exhaust gas temperatures are also obtained, which does not allow obtaining high potential steam in the steam generator. All these signs have a significant impact on the implementation in diesel engines of spraying water in the intake air, as well as obtaining high potential steam for the job.

When water is added to the intake air, a decrease in the temperature and pressure of the end of compression occurs, which leads to an increase in the ignition delay period and, consequently, the rigidity of the engine. In this case, it is necessary to inject so much water into the air and so cool it in the air cooler so that P _c and T _c are acceptable for soft operation of the engine. In naturally aspirated engines, P _c and T _c are maintained by recirculation of the combustion products.

The low temperature of the exhaust gases in these engines does not allow steam with a high temperature to be obtained even at the nominal operating mode, and at partial loads this temperature is further reduced, and therefore, the efficiency of using the resulting steam and water is also reduced. An increase in the temperature of the exhaust gases in nominal mode was achieved by reducing α to 1.1, and maintaining this α and maintaining a high temperature of the combustion products at partial loads was achieved due to quantitative power control. In this case, reliable ignition of the fuel at these loads is carried out by recirculation of the combustion products or by reducing the cooling of the charge air.

With such an organization of the working process in adiabatic diesel engines, there will also be a significant increase in their effective efficiency and power, not only due to an increase in the number of working fluid performing work, but also due to the use of high-potential steam to perform work in the converted steam (s) or gas cylinders.

The device of a combined engine with a two-phase working fluid contains a reciprocating internal combustion engine 1 (Fig. 1) with heat-insulated parts of the combustion chamber [cylinder head (caps), pistons and gas cylinders 2], which are connected in series with the boiler 3, combined with a thermally insulated exhaust manifold 4 , a condenser 5 and a turboexpander (turbine) 6 by means of a gas pipeline 7. For air intake into these cylinders, a compressor 8 is arranged in series, connected to a turboexpander (turbine) 6; the intake manifold 9 is connected to the compressor 8 by an air duct 10. To recover the heat of exhaust gases and oil, this engine, in addition to boiler 3, is also equipped with a hot water distribution device 11, divided by partitions 12 and 13 into upper 14, middle 15 and lower 16 parts. In the middle part 15, a water-oil heat exchanger 17 is located. The water from the combustion products is regenerated by a condenser 5 together with a cold water distribution device 18, which are connected by a water supply 19, with a water pump 20 installed on it and a flow regulator 21. For quasi-isothermal compression of air by compressor 8, it is cooled produced by spraying cold water with a sprayer 22 connected to a cold water distribution device 18 by pipelines 19 and 23, the last of which is installed the flow regulator of this water is connected 24. The condenser 5 is connected to the middle 15 and lower 16 parts of the hot water distribution device 11 by the water pipelines 25, 26 and 27, a water filter 28 is installed on the water supply 25, and a flow regulator 29 is installed on the water supply 27. The atomizer 33 of the condenser 5 is connected to the water supply system 19 through the flow regulator 21. The boiler 3 is connected to the hot water distribution device 11 by the water supply 30, on which the feed pump 31 and the flow controller 32 are located. Electromagnetic nozzles 34 for supplying water to the cylinders are installed They are located in the heat-insulated (s) cylinder head (s) 2 and are connected to the boiler 3 with heat-insulated heat pipes 35, 36 and 37. The nozzle holes of the nozzles in the cylinder head (s) are installed tangentially or at an acute angle to the walls of the combustion chamber so that the resulting water the steam at the exit from them into the cylinders swirled the combustible mixture. To supply fuel to the engine 1, electromagnetic nozzles 38 are installed on the inlet manifold 9 nozzles 38, which are connected to the fuel pump 39 through the fuel filter 40 and the flow regulator 41 by the fuel lines 42 and 43. The heat-insulated water distribution devices of hot 11 and cold 18 water are connected to each other by a water pipe 44 and contain air 45 and steam 46 valves. In a stationary engine, in order to maintain the water level in the water distribution devices, one of these devices is connected to a water pipe 47 on which a water level regulator 48 and a shut-off device 49 are installed. The nozzle guide apparatus 50 of the turboexpander (turbine) 6 changes the pressure in the condenser 5 and the pressure of the blower air compressor 8. The control panel 51 is used to control the operation of the engine in manual and automatic modes, controls the correct operation of the engine and protects it from damage.

To supply steam to the thermally insulated cylinders 2 (FIG. 2), steam inlet valves 52 with swirls were used instead of nozzles to improve mixture formation in the combustion chamber. Steam is supplied to these valves from boiler 3 via heat lines 35, 36, and 37, like water. To regulate the amount of steam supplied, a flow regulator 32 is located on the water supply pipe 30, nozzles 38 for spraying fuel are installed on the inlet pipes of the intake manifold 9. To supply fuel to them, fuel lines 43 and 42 are laid, the last of which has a fuel pump 39, a fuel filter 40 and flow regulator 41. In this diagram, there is no turbo expander (turbine) and compressor in order to simplify the design of the engine. To change the pressure in the condenser 5, a throttle valve 53 mounted on the exhaust gas line 54 is used. Air in the heat-insulated cylinders 2 of both gas and gasoline (with positive ignition) and diesel (self-igniting) engines is supplied through the intake manifold 9 through the throttle valve 55, installed on the air duct 56, to change the load of the engines at a constant coefficient of excess air.

The diagram in FIG. 3 differs from the previous schemes in that for converting the steam produced in the boiler 3 into operation, one of several thermally insulated gas cylinders 2 is converted into steam 57. In this scheme, the boiler 3 is connected to the inlet steam valve 58 by a steam line 59, and the exhaust steam a valve 60 of this cylinder is connected to the condenser 5 through a non-return valve 61 by a steam line 62. The boiler 3 is connected to the exhaust heat-insulated gas cylinders 2 through the nozzles of the exhaust manifold 4, and to the condenser 5 by a discharge gas line 7, s which is turboexpander (turbine) 6 connected to the compressor 8. This compressor is connected to an inlet collector 9 duct 10. Nozzles 22 and 33 are mounted for spraying water in the air and in the flue gas entering the compressor to the condenser 8 and 5, respectively. A nozzle guide apparatus 50 in front of the turboexpander (turbine) 6 is installed to control the required pressure in the condenser 5, and therefore the condensate temperature, depending on the outdoor temperature. The condenser 5 is connected to the hot water distribution device 11 by the water supply lines 25, 26 and 27. A water filter 28 is installed on the water supply pipe 25, and a flow regulator 29 is installed on the water supply pipe 27. For cooling the condensate, a closed circulation circuit is organized in the cold water distribution device 17, which includes water pipes 63 and 64, a water-air radiator 65 and a water pump 66 with a flow regulator 67. This radiator is cooled by a fan 68. A heat-insulated tank 69 for draining water from this system is included in the water and steam supply system when it freezes, which contains an electric heater 70, air 71 and steam 72 valves and a filler neck with plug 73. It is connected to hot water 11 and cold 17 water distribution devices with water pipes 44 and 74, the last of which has a shut-off device 75. To fill the system An electric pump 76 is installed on the water supply 77. A heat-insulated tank 69, after draining the water from the system, can be disconnected and transferred to a warm room. Before connecting this tank to the water and steam supply system, the water in it can be heated by an electric heater 70 to the boiling temperature from the mains.

In the diagram of Fig. 4, in contrast to the diagram in Fig. 1, nozzles 77 are shown in the inlet nozzles of the manifold 9 for spraying cold water in the air entering the gas insulated gas cylinders 2. These nozzles are connected by pipelines 78, 79 and 80 to a cold water distribution device 17 A water pump 20 is installed on the water supply 78, and a water flow regulator 24 is installed on the water supply 79. A carburetor 81 is installed to spray the fuel in the air entering the cylinders through the intake manifold 9 and is connected to the fuel pump 39 82, on which the fuel filter 40. In this diagram, no expansion turbine (turbine) and a compressor, although they could also be connected to the engine and this engine power is only increased.

In the diagram of Fig. 5, for supplying a fuel-water solution to the air entering the heat-insulated cylinders, a carburetor 83 is installed in which this solution is formed. To supply fuel to this carburetor, it is connected to a fuel line 82, on which a fuel pump 39, a fuel filter 40 and a flow regulator 41 are installed. For water supply, it is connected to water pipes 78 and 79. A pump 20 is installed on the water supply 78, and 79 on the water supply water flow regulator 24.

To prepare the fuel-water mixture, the engine is equipped with a mixer 84 (6), which is connected to the fuel pump 39 through the fuel filter 40 with a fuel pipe 82, on which the fuel flow regulator 41 is mounted. This mixer is connected to the cold water distribution device 17 with water pipes 78 and 79 moreover, a water pump 20 is installed on the water supply 78, and a water flow regulator 24 is installed on the water supply 79. For supplying the fuel-water mixture to the nozzles 85 of the gasoline (with forced ignition) engine located in pat cuttings of the intake manifold 9, they are connected to the mixer 84 by a pipe 86, on which a pump 87 for supplying this mixture and a flow regulator 88 are installed. To supply this mixture with a diesel (self-igniting) engine, nozzles 89 are located in the insulated cylinder head (s) and are connected to the mixer 84 by a pipe 90 on which a high pressure pump 91 and a flow regulator 92 with a pressure reducing valve 93 are installed. To control the operation of the engine, it is equipped with a control panel 51.

For double or multiple circulation of feed water through the boiler at a temperature of combustion products behind the boiler above 150 ° C, a closed circulation circuit (Fig. 7) is created, which includes the feed pump 31, boiler 3, the first shut-off valve 94 and the heat pipes 95 and 96. When In this case, the reverse supply of water to the hot water distribution device 11 is excluded by the non-return valve 97, and the supply of steam to the steam cylinder 57 is interrupted by the second shut-off valve 98 installed on the steam line 59. The disconnected steam heat-insulated cylinder 57 is connected It is connected to the atmosphere through a steam-air valve 99 to reduce the operation of pump strokes. All these valves 94, 98 and 99 are connected to the logic device 100, and the logic device is connected to the control panel and the exhaust gas temperature sensor 101.

To mix the water in hot and cold water distribution devices, they are designed in a cylindrical shape (Fig. 8), and the water is supplied to them through tangentially arranged pipes 102. This water is taken from these devices through perforated pipes 103 located vertically along the axis of these distribution devices. , the ends of which go out through the upper 104, lower 105 or side 106 of the wall.

The method of operation of the combined engine with a two-phase working fluid is carried out by supplying air to the compressor 8 (Fig. 1) in heat-insulated gas cylinders 2 through the air duct 10 and the intake manifold 9. The air is compressed in the compressor 8 almost without increasing its temperature due to atomization of cold water by the atomizer 22 in the intake the nozzle of this compressor in the amount of 0.5 ÷ 0.75 from the cyclic fuel supply with droplet sizes such that the effective efficiency of the cycle becomes greatest, and detonation does not occur. At the same time, the maximum temperature of the cycle is significantly reduced, which allows to increase the degree of compression and to further increase this efficiency. Water is sent to the atomizer 22 by the pump 20 through the water supply system 19 from the cold water distribution device 18, while the amount of water supplied is regulated by the flow regulator 24, and the size of the droplets is determined by the design of the atomizer 22. When compressed in the compressor 8 and in the heat-insulated gas cylinders 2, they turn into a heat carrier with a temperature close to the boiling point at the end of compression, when heat is supplied, pass into the working fluid, significantly reducing the temperature of the gases and eliminating detonation, as it expands would be formed an additional amount of the working fluid performs work by increasing the efficiency of the engine with a marked reduction in maximum combustion temperature. Thanks to this, not only does knocking not occur, but the heat stress of both gasoline (with forced ignition) and diesel (with self-ignition) engines is also significantly reduced.

In gasoline (forced ignition) engines, there is no danger that due to air cooling with water, misfire may occur in the mixture because the ignition temperature of gasoline-air mixtures is significantly lower than the temperature of this mixture at the time of ignition. In diesel (self-igniting) engines, such a danger exists at the time of starting the engine, and after entering the mode and warming up the engine, it also disappears, but the engine starts working harder than the base one due to an increase in the ignition delay period and an increase in the dynamic factor of the cycle. To avoid this negative phenomenon that occurs when water is injected into the intake air, it is necessary to further increase the compression ratio of this engine, which is already high. Another way is to supply heated water in the boiler 3 through the plumbing 35, 36 and 37 to the nozzles 34 located in the heat insulated and to the cylinder head (caps) at the beginning of the expansion, while the injection pressure exceeds the maximum combustion pressure and the water temperature equals the saturated temperature steam at this pressure and the amount of water supplied is regulated by a flow regulator 32 according to the maximum effective diesel efficiency. With such a supply of water to the nozzles, not only does this efficiency increase, but the maximum combustion temperature also decreases, which can significantly reduce the thermal stress of this engine.

The combustion products from the heat-insulated gas cylinders 2 when opening the exhaust valves are sent to the boiler 3, combined with a heat-insulated exhaust manifold, where they heat water or generate steam. Then, the combustion products through the gas pipeline 7 enter the condenser 5 for cooling to the temperature of condensation of water vapor, which depends on the pressure in this condenser. This pressure is controlled by changing the flow area of the nozzle guide apparatus of the turbo expander (turbine) 50 so that the condensate formed at this pressure can be cooled in the cold water distribution device 17 due to its circulation through the air-water radiator 65 (Fig. 3) by the water pump 66 and under the action fan 68.

The temperature of the exhaust gases coming from the heat-insulated exhaust manifold to the condenser depends on the engine working process and, first of all, on its coefficient of excess air in the cylinders, the amount of coolant passing through the boiler, its temperature and pressure, the diameter of the tubes and their heat transfer surfaces, and also the gas velocity relative to the heat-absorbing surfaces of the boiler. The task is to ensure that the temperature of these gases does not exceed 150 ° C, and the enthalpy of the generated steam or water is greatest. The temperature of the flue gases from this collector is measured by a thermometer, and the enthalpy is determined by the measured pressure, temperature and flow rate of the coolant and is calculated according to tables or empirical formulas. With the constant structural dimensions of the boiler, the maximum enthalpy of the coolant is regulated by the flow regulator 32 (Fig. 1) and by changing the pressure of the feed pump 31 using its pressure reducing valve. The lower the temperature of the exhaust gases from the boiler, the smaller the loss of heat with the exhaust gases, and therefore the size of the devices involved in the condensation of water vapor, as well as the cooling of water by air. Calculations showed that for a transport engine this temperature should not exceed 150 ° C, since in this case it is possible to preserve the design and dimensions of the air-water radiator of the base engine.

In gasoline (with forced ignition) and diesel (self-ignition) engines, it is necessary to maintain such a low coefficient of excess air in all operating modes in order to have the highest temperature of the exhaust gases in order to obtain a high-potential coolant for water or steam. For gasoline (with forced ignition), this coefficient of excess air should be α≅1.0 ÷ 1.2, and in diesel (with self-ignition) α≥1.1. For other values of α <1 and α> 1.2 in gasoline (with forced ignition) engines, the coefficient of active heat generation decreases, and in diesel (with self-ignition) at α <1.1, the temperature of the exhaust gases decreases and additional work is spent on supplying air to cylinders. The coefficient of excess air in the cylinders is determined at the stage of creating a prototype or industrial sample by measuring the amount of oxygen in the combustion products, and is also regulated by the amount of fuel supplied to the cylinders by the fuel consumption regulator 41.

In traditional diesel (with forced ignition) (self-ignition) engines, when switching to partial loads, the coefficient of excess air rapidly increases due to the quality control of its power, which leads to a significant decrease in the temperature of exhaust gases and a significant decrease in both the temperature of the water in the boiler and steam if it will be produced. In order for this system to work efficiently in a diesel (self-igniting) engine and at partial loads, it is necessary to introduce a quantitative power control such as in a gasoline (forced ignition) engine. The throttle valve 55 (figure 2) when switching to partial loads reduces the amount of incoming air into the insulated cylinders 2, and the fuel flow regulator 41 together with the nozzles 38 reduce the amount of incoming fuel into the insulated cylinders 2, and the excess air ratio of more than 1.1 is regulated by these devices together depending on the quality of fuel combustion, by the absence of soot content in the combustion products.

The design of a combined engine with a two-phase working fluid can be simplified if the turboexpander (turbine) is replaced by a throttle valve 53 (Fig. 2), and the compressor is excluded from the design. Air is supplied to the cylinders 2 by suction through the intake manifold 9. At the same time, the pressure in the condenser 5 can be controlled by changing the position of the throttle valve 53, which will lead to a noticeable increase in the operation of pump strokes and a decrease in the effective efficiency depending on the degree of covering of this valve.

Steam is supplied to heat-insulated cylinders 2 through inlet steam valves 52 (Fig. 2) through steam lines 35, 36 and 37 from boiler 3. In this case, the steam pressure in the steam lines is provided by the feed pump 31, and the amount of steam supplied is regulated by the maximum effective engine efficiency and supported by flow control 32.

The cylinder of a traditional piston engine with heat-insulated cylinder head (s) and gas cylinders is converted into a steam heat-insulated cylinder 57 (Fig. 3) with inlet 58 and outlet 60 steam valves. Steam is supplied to the steam cylinder 57 through the inlet steam valve 58 when it is opened at the top (inner) dead center, and it is closed at the moment of permissible reduction of the vapor pressure in the boiler, at which the effective engine efficiency begins to decrease. The steam exhaust valve is opened when the steam pressure in the steam cylinder still exceeds the pressure in the condenser, and steam begins to flow through the steam line 62 through the non-return valve 61 to the condenser 5. The non-return valve 61 prevents gases from the condenser 5 from flowing back into the steam cylinder 57 when gases are released from other cylinders.

The easiest way to control the amount of feed water to the boiler 3 by the controller 32 is according to the required temperature and steam pressure in the boiler or the temperature of the exhaust gases behind the boiler, entering the condenser 5.

When the combined engine is operating, the required water level 12 in the water distribution devices of hot 11 and cold 17 water during operation at various operating modes is kept constant. If this level begins to decrease, then the guide nozzle device 50 in the turboexpander (turbine) 5 increases the pressure in the condenser 5 at a constant temperature of the gas-vapor mixture and thereby increase the amount of condensate formed. If the water level in these water distribution devices begins to rise, then the pressure in the condenser is reduced so as not to increase pump losses and not to reduce the effective engine efficiency. In stationary engines, the flow area of the turbine (turboexpander) is not changed in order not to reduce the effective efficiency of the engine, and when the amount of water in the steam-water system decreases, it is automatically replenished from the water supply using the flow regulator 48 (Fig. 1).

To cool the combustion products entering the condenser 5 (figure 3), cold water is required. Therefore, the condensate formed is cooled in the cold water distribution device 17 by circulating it by a pump 66 through a water-air radiator 65, through water pipes 63 and 64, the last of which has a flow regulator 67. Fan 68 controls the amount of heat transferred to the outside air depending on the engine operating mode and outdoor temperature. The flow regulator 67 and fan 68 together provide a heat transfer coefficient at which the temperature of the cold water is sufficient to cool the combustion products in the condenser to the condensation temperature of water vapor, at which there is no loss of water in the steam-water system. At lower outdoor temperatures, the power consumption of the pump and fan is noticeably reduced, and the effective motor efficiency is increased.

When the engine stops and the outside temperature is low, there is a danger of freezing of water in the steam-water system. When the water temperature drops close to the freezing temperature, a signal is sent from the temperature sensor to the control panel and the engine starts to warm up manually or automatically. Another option to protect the steam-water system from freezing is to drain the water from it manually or automatically by command from the control panel 51 (Fig. 3) into a heat-insulated container 69. For this, air valves 45 in the water distribution devices 11 and 17 and the steam valve 72 in this tanks, and also the shutoff valve 75 from the control panel. After draining the water, the insulated container 69 is disconnected from the engine and transferred to a warm room. In this room, the water in it is heated before starting the electric heater 70 to a boiling point. Then it is connected to the engine and the system is charged with electric pump 76 from the control panel with shut-off valve 75 closed, and this pump is turned off. This excludes not only the freezing of the steam-water system, but also the cold engine warms up before starting, which increases the vehicle mileage by more than 100 km compared to starting a cold engine.

Both in gasoline (with forced ignition) and diesel (with self-ignition) engines, a short-term increase in power is carried out by increasing the amount of combustible mixture or air entering the cylinders. In these engines, such an increase in power is achieved with a constant coefficient of excess air α in gasoline (with positive ignition) engines α≥1.0, and in diesel (with self-ignition) α> 1.1 by a sharp increase in the pressure of the charge air by compressor 8 (Fig. 1) due to the increase in gas pressure in front of the turboexpander (turbine) as a result of a decrease in the orifice of the nozzle guide apparatus 50. The amount of harmful emissions in the combustion products does not increase, the effective efficiency increases While increasing mechanical efficiency, and complete combustion of fuel is maintained.

The supply of fuel and water to the cylinders with an electronic fuel management system complicates the design of the engine. Therefore, it is advisable to saw the fuel in the intake air with a carburetor 81 (Fig. 4), and saw the water in the air in the intake pipes with the nozzles 77 and pump it through the water pipes 78, 79 and 80, while the amount of water supplied by the nozzles 77 is regulated by the flow regulator 24. The design of the nozzles of these nozzles provides such a spraying of water in the air, in which the droplets formed during compression become heat carriers and by the end of the compression are heated to boiling point, when heat is supplied, they turn into e rs body and during the expansion of the increased amount of the working fluid performs extra work. In this case, the gas temperature, according to the studies conducted by the authors, significantly decreases despite the flow of the working process in the insulated space of the combustion chamber, which allows to increase the compression ratio of the engine and thereby increase its effective efficiency even more. If water is sprayed with nozzles 77 in each inlet, then knocking can be suppressed by controlling the amount of water supplied. The studies conducted by the authors showed that changing the optimal amount of water supplied has a weak effect on reducing the effective efficiency of the engine, but can significantly reduce the requirements for the octane number of fuel. As a result, instead of the traditional system of changing the ignition timing, to suppress detonation, a water injection system is used by nozzles in the inlet pipes or heat-insulated gas cylinders 2, which increases the fuel economy of the engine in operating conditions.

The design of the engine can be simplified by supplying fuel and water together, after mixing them. To mix these two liquids, the carburetor of the base engine can be used or a special mixer can be made. In Fig. 5, fuel is supplied through the fuel line 82 by the fuel pump 39 through the fuel filter 40 to the carburetor 83. The amount of fuel supplied is regulated by the flow regulator 41. Water to the carburetor 83 is pumped by the pump 20 through the water pipes 78 and 79. The amount of water supplied is regulated by the flow regulator 24 and is 0.5 ÷ 0.75 of the cyclic fuel supply. To eliminate detonation of the engine by adjusting the amount of sprayed water, the cyclic water supply in some modes, depending on environmental conditions, can increase to 1.25 from the cyclic fuel supply, while the heat utilization factors in bm υ _z and at the end of the expansion υ _b do not decrease. Spray quality (average droplet size) is determined experimentally on a pilot or industrial sample by the highest effective efficiency when changing the diameters of the diffuser and carburetor atomizer. Mixing water and fuel in the float chamber is carried out by any of the known methods.

In the diagram (Fig. 6), the mixing of fuel and water is carried out in a cylindrical mixer 84 by tangential supply of fuel through the fuel pipe 82 by pump 39 through the fuel filter 40. The amount of fuel supplied is regulated by the flow regulator 41. Water is supplied to this mixer also tangentially through the water pipes 78 and 79 pump 20. The amount of water supplied is regulated by the regulator 24 and is 0.5 ÷ 0.75 of the cyclic fuel supply in normal operation modes, and when eliminating detonation, it can reach 1.25 of the cyclic fuel supply. From the mixer 84, the fuel-water mixture is supplied by the pump 87 through the pipeline 86 to the nozzles 85 located on the inlet pipes of the manifold 9 of the gasoline (forced ignition) engine. In a diesel (self-igniting) engine, this mixture is fed to nozzles 89 located in the heat-insulated cylinder head (s) through pipe 90 by pump 92. The amount of the supplied mixture is regulated by nozzles 85 and a flow regulator 88 in gasoline (with forced ignition) engine, and in a diesel (self-igniting) engine, nozzles 89 and a flow regulator 92.

When the temperature of the exhaust gases behind the boiler drops below 150 ° C, when the enthalpy of steam is greatly reduced, the water and steam circuits are turned on through the boiler 3 (Fig. 7) under the action of the pump 31 through pipelines 95 and 96. At this point, a signal is received from the temperature sensor to the control panel, and from there to the logic device 100, which instructs to open the first shutoff valve 94 and close the second shutoff valve 98. In this case, the check valve 97 and the second shutoff valve 98 are closed. The supply of steam to the thermally insulated steam cylinder 57 ceases, and the steam begins to circulate through the boiler and the first open shut-off valve 94 through lines 95 and 96 until the temperature of the exhaust gases rises above 150 ° C. During this period, the open steam-air valve 99, upon command from the logic device 100, informs the steam insulated cylinder 57 with the atmosphere and thereby reduces the operation of the pumping strokes of the steam cylinder to zero. After warming up the steam to the required temperature and increasing the temperature of the exhaust gases behind the steam generator above 150 ° C, the steam valve 99 and the first shut-off valve 94 immediately close, the check valve 97 and the second shut-off valve 98 open. High-potential steam from boiler 3 starts to flow through the steam line 59 to the steam heat-insulated cylinder 57 and produce work there.

To mix heat carriers with different temperatures in the used cylindrical devices of the combined engine, as well as to increase the heat transfer coefficient from oil to water through the walls of the oil-water heat exchanger in the water distribution device for hot water, tangential coolants were supplied through pipes 102 to these devices (Fig. 8), and to pick up mixed coolant from these devices, it must be removed through perforated cylindrical pipelines 103 located in the center of this device. The end of this pipeline can be led out through the bottom 105, top 104 or side 106 of the walls of these tanks.

The design and method of increasing the triggered heat drop on the impeller of the turboexpander are described in 1 and 2 (see below).

The resulting detonation is suppressed in the same way as in the manufactured transport gasoline engines of OJSC “ZMZ” (see 3 on page 69), only here is the change in the amount of added water in the air instead of changing the ignition timing.

The water supply to the boiler is implemented in the same way as the fuel supply in diesel engines and water in heat generating units (see 4 and 5 on page 69).

The amount of steam supplied to the gas cylinders is determined by conducting experiments on an industrial or experimental sample with various degrees of compression and quantities of steam supplied. Their rational or most advantageous values at various operating modes are determined by the absence of detonation and the highest effective efficiency, which is compared with the effective efficiency of the base engine.

The designs of shut-off devices for draining water from a two-phase water and steam supply system can be designed similar to those in 3 and 5 (see page 69), where they are used to drain fuel, oil, coolant, coolant, condensate, etc. . in piston engines and thermal power plants.

The sizes of nozzles and pump nozzles with the calculated average droplet diameter according to formula 38 and the known pump flow are determined in the same way as when creating a fuel supply system for a new diesel engine (see below 4).

The dimensions of the carburetor atomizer and its diffuser with the calculated average droplet diameter according to formula 38 are determined in the same way as when creating a carburetor for a gasoline engine (see below 6).

The sizes of nozzle nozzles are calculated by the average droplet diameter determined by formula 38, and the dimensions of the pump by the known amount of water supplied or a mixture of water and fuel (see below 4).

1. Stepanets A.A. Energy-saving turboexpander units. Nedra, 1999.

2. Davydov A.B., Kabulashvili A.Sh., Sherstyuk A.N. Calculation and design of turbo expanders. - M.: Mechanical Engineering, 1987.

3. Engines ZMZ-406 GAZ and UAZ vehicles. Design features. Diagnostics. Maintenance. Repair / Giryavets A.K., Golubev P.A., Kuznetsov Yu.M. et al. - Nizhny Novgorod: NSU im. Lobachevsky, 2000.

4. Supply and atomization of fuel in diesel engines. Astakhov N.V., Trusov V.I., Khachiyan A.S. et al. - M.: Mechanical Engineering, 1971.

5. Reznikov M.I., Lipov Yu.M. Boiler plants for power plants. - M: Energoatomizdat, 1987.

6. Car engines. Arkhangelsky V.M., Wikhert M.M., Voinov A.N. et al. - M.: Mechanical Engineering, 1967.

The advantages of the developed method of operation and the device of a combined engine with a two-phase working fluid based on a reciprocating internal combustion engine are as follows: the design and method of operation of the combined engine are developed not only on the basis of gasoline and gas (with positive ignition), but also on the basis of diesel (with self-ignition ) engines with a two-phase working fluid, the effective efficiency of which increases to the extent limited by the first and second laws of thermodynamics ≈0.6; the design of this engine is simplified due to the atomization of water not only in the inlet pipe of the compressor, but also in its supply or supply of steam to the combustion chambers during the expansion period in thermally insulated gas cylinders; the complexity of manufacturing and the cost of the engine are reduced as a result of using one fuel-water system instead of two water and fuel supply systems; this engine can be a simpler design without a compressor and a turboexpander (turbine) with only less effective power, but with a higher effective efficiency than the base engine, due to the use of a throttle valve in the exhaust gas pipeline after the condenser; the developed engine is more reliable during its operation in conditions of low outdoor temperatures, at which there is a possibility of freezing of the steam-water mixture; a short-term increase in power (throttle response) is not lower than that of a basic gasoline (with forced ignition) engine with its higher efficiency and environmental safety.

Claims (31)

1. The method of operation of a combined engine with a two-phase working fluid based on a reciprocating internal combustion engine, including supplying air to the intake manifold by a compressor, cooling it by spraying with cold water, cooling the combustion products in series in the boiler and in the contact condenser to a temperature at which the condensate released can be cooled to cold water temperature with outside air in a water-air radiator, then their expansion in a turboexpander (turbine), separation with a drop of water in a contact cond when expanding the combustion products, supplying cold water to the contact condenser for cooling the combustion products and to the intake system for cooling the air, reducing the maximum combustion temperature due to the evaporation of water droplets when burning fuel in the combustion chambers, heat transfer from the oil of the lubrication system heated water in the water distribution device through the dividing walls of the oil-water heat exchanger, the supply of heated water to the boiler from the water distribution device and steam supply from the boiler to an expansion device to obtain useful work, regulating the gas pressure in the condenser by throttling it in front of the impeller of the turboexpander (turbine) depending on the outdoor temperature, characterized in that they reduce the maximum combustion temperature due to the evaporation of water droplets during the combustion of fuel in the combustion chambers at a coefficient of excess air equal to 1.0 ÷ 1.2 in gas and gasoline (with forced ignition) and the lowest, but greater than 1.1 in diesel (with self-ignition) engines fir trees. 2. The method according to claim 1, characterized in that the water is supplied to the air entering the compressor of a gasoline (forced ignition) engine to reduce the temperature of this air when it is compressed in the compressor. 3. The method according to claim 1, characterized in that in a diesel (self-igniting) engine, heated water is supplied to the boiler to a boiling point at a pressure higher than the maximum combustion pressure, in gas insulated gas cylinders at the beginning of expansion when burning fuel with the lowest excess ratio air, but large 1.1. 4. The method according to claim 3, characterized in that the water is sprayed in the heated air when it is compressed in a compressor or in cylinders. 5. The method according to claim 1, characterized in that the flow area of the guide nozzle apparatus of the turboexpander (turbine) is reduced while maintaining the coefficient of excess air in the cylinders α equal to 1.0 ÷ 1.2 in gas and gasoline (with forced ignition) and α large 1.1 in diesel (self-igniting) engines. 6. The method according to claim 2, characterized in that in gas and gasoline (with forced ignition) engines suppress the detonation that occurs by increasing the amount of cold water sprayed in the air entering the cylinders. 7. The method according to claim 1, characterized in that the feed water is circulated twice or repeatedly through the boiler at a temperature of combustion products behind the boiler above 150 ° C through a closed circulation circuit with shut-off valves that are connected to a logic device connected to the control panel and the sensor exhaust temperature. 8. The method according to claim 1, characterized in that the generated steam in the boiler is fed into the gas cylinders at the beginning of the expansion through the inlet steam valves and swirls. 9. The method according to claim 7, characterized in that the amount of feed water to the boiler is controlled by the temperature of the exhaust gases at the outlet of this boiler. 10. The method according to claim 7, characterized in that the amount of feed water to the boiler is regulated by the pressure in it of steam. 11. The method according to claim 8, characterized in that the amount of feed water to the boiler is controlled by the temperature of the steam in it. 12. The method according to claim 1, characterized in that the mixture of gasoline and water is sprayed in a ratio of 0.5 ÷ 1.25 from the cyclic supply of fuel in the air entering the cylinders of a gasoline engine. 13. The method according to claim 1, characterized in that the mixture of water and diesel fuel is sprayed in the air, while providing a water supply with a flow regulator of 0.5 ÷ 1.25 from the cyclic supply of fuel in the air entering the cylinders of the diesel engine. 14. The method according to claim 1, characterized in that the steam is supplied to the steam (s) cylinder (s) through the inlet (s) of the steam (s) valve (s) at its (their) opening at the top (inner) dead point. 15. The method according to claim 1, characterized in that when the water level in the water distribution devices of cold or hot water decreases, the gas pressure in the condenser is increased or water is added to these tanks. 16. The method according to claim 1, characterized in that when the engine is off and the temperature of the cold water drops to freezing temperature, the engine starts automatically or manually. 17. The method according to claim 1, characterized in that when lowering the temperature of cold water in the engine to freezing temperature, water from the water and steam supply system is automatically or manually drained through a pipe into an insulated tank with an electric pump, and when it is started, it is heated with an electric heater and refilled into this system from this tank. 18. The method of operation of a combined internal combustion engine with a two-phase working fluid, comprising supplying air, cooling it by spraying cold water in an amount of 0.5 ÷ 0.75 from a cyclic fuel supply, cooling the combustion products sequentially in the boiler and in the contact condenser to a temperature, in which the condensate which has evolved is cooled to the temperature of cold water with outside air in an air radiator, the supply of cold water to the contact condenser for cooling the combustion products and to the intake system for cooling air, lowering the maximum temperature of combustion due to the evaporation of water droplets when burning fuel in the combustion chambers, the transfer of heat from the oil of the heated water lubrication system in the hot water distribution device through the separating walls of the oil-water heat exchanger, the supply of heated water to the boiler from this water distribution device and the supply of steam from the boiler in the expansion device to obtain useful work, characterized in that the air is supplied to the cylinders by suction through the intake manifold, the pressure in the condenser is controlled by changing the position of the throttle valve, lowering the maximum combustion temperature due to the evaporation of water droplets when burning fuel in the combustion chambers with an excess air coefficient of 1.0 ÷ 1.2 in gas and gasoline (with forced ignition) and the lowest , but greater than 1.1 in diesel (self-igniting) engines. 19. A device of a combined engine with a two-phase working fluid, containing gas cylinders of a reciprocating engine for expanding combustion products, a boiler combined with an exhaust manifold, a condenser with a turboexpander (turbine) located in series along the exhaust gas pipeline and connected by gas communication with the cylinders of this engine, a compressor, connected to cylinders for air supply and connected to a turboexpander (turbine), as well as hot and cold water distribution devices connected to waiting for myself, with a condenser and a water-air radiator by pipelines, fuel nozzles in the inlet pipes, connected by pipelines to a fuel pump, filter and flow regulator for supplying fuel to the cylinders, characterized in that it contains heat-insulated parts of the combustion chamber (cylinder head (covers), pistons and gas cylinders), a boiler combined with a thermally insulated exhaust manifold, nozzles and a high pressure pump with a flow regulator for supplying water to the cylinders from the boiler at the beginning of expansion, while p organizes swirling motion of the combustible mixture in the combustion chamber to improve mixing and combustion processes. 20. The device according to claim 19, characterized in that it contains a thermally insulated container with high resistance to heat transfer from the walls, an electric water heater, air and steam valves, a filler neck, which is connected through an electric pump and a shut-off valve to the water and steam supply system, this electric pump is connected with remote control, and this capacity can be disconnected from this system and transferred to a warm room for storage, as well as heating water in it to heat the engine before starting. 21. The device according to claim 19, characterized in that it contains thermally insulated water distribution devices for hot and cold water and shut-off valves. 22. The device according to claim 19, characterized in that it contains a hot water distribution device divided by perforated partitions into three parts, in the middle part of which there is a water-oil heat exchanger, condensate supply pipes from the condenser and heated feed water to the boiler, in its lower part there are pipes for supplying condensate from a condenser and for discharging mixed water into a cold water distribution device. 23. The device according to claim 19, characterized in that it contains nozzles in the inlet pipes connected to the water pipes for supplying cold water from the water distribution device, on which the pump and the flow regulator are installed, and also has a carburetor connected to the intake manifold and the fuel pump by a pipeline on which the fuel filter is located, while the pump provides water supply along with a flow regulator 0.5 ÷ 1.25 from the cyclic fuel supply. 24. The device according to claim 19, characterized in that it contains a carburetor connected to the intake manifold, a fuel line for supplying fuel with a fuel pump through a fuel filter and a fuel flow regulator, and a cold water supply for supplying it with a water pump through a water flow regulator in an amount 0.5 ÷ 1.25 from the cyclic supply of fuel, while the fuel is mixed with water in the float chamber. 25. The device according to claim 19, characterized in that it contains a fuel and water mixer, which is connected to the cold water distribution device by two water pipes, a water pump is installed on the first one, and a flow regulator on the second one, it is also connected to the fuel pump by a fuel pipe on which the fuel filter and the fuel consumption regulator are located, in addition, this mixer is connected by pipelines through the pump and the flow regulator to the nozzles of the gasoline (forced ignition) engine located on inlets of the intake manifold, or it is connected by a pipe through a high-pressure pump, a flow regulator and a pressure reducing valve with diesel engine nozzles (self-igniting) installed in the insulated cylinder head (s), while the water pump together with the flow regulator provide water supply 0.5 ÷ 1.25 from cyclic fuel supply. 26. The device according to claim 19, characterized in that a closed circulation circuit is formed under the action of the feed pump through the boiler and the first locking device associated with the logical telecommunication device, for normal operation on the steam line connecting the boiler to the steam inlet (s) ( i) the valve (s) of the steam (s) cylinder (s) is equipped with a second locking device and a valve for communicating this (their) steam (s) cylinder (s) with the atmosphere to reduce the operation of the pumping passages of the disconnected (s) steam (s) cylinder (s) that are bonded Ana telecommunications with a logical device. 27. The device according to claim 19, characterized in that in the cylindrical devices used, the inlet and outlet pipes of cylindrical liquids are located tangentially to the inner cylindrical surfaces of the containers for mixing heat carriers with different temperatures, as well as increasing the heat transfer coefficient through the walls of the oil-water heat exchanger, while the pipes water drains are located vertically in the center of these devices, and have perforated holes for the intake of mixed water, and they exit through the sides e upper or lower wall outwardly. 28. The device according to claim 19, characterized in that it contains a water pump and nozzles (sprayers) for adding water to the air in the inlet pipes of gas and gasoline (with forced ignition) engines, a water pump and nozzles for supplying water to insulated diesel cylinders ( with self-ignition) of the engine and spraying it there in the air, while these pumps provide a water supply of 0.5 ÷ 1.25 from the cyclic fuel supply. 29. The device according to claim 19, characterized in that it contains a control panel for engine operation in manual and automatic modes, which is associated with temperature, pressure, detonation and flow rate sensors, as well as logic and actuators. 30. Device of a combined engine with a two-phase working fluid, containing gas cylinders of a reciprocating engine for expanding combustion products, a boiler combined with an exhaust manifold, a condenser with a turboexpander (turbine) arranged in series along the exhaust gas pipeline and connected by gas communication with the cylinders of this engine, a compressor, connected to cylinders for air supply and connected to a turboexpander (turbine), as well as hot and cold water distribution devices connected to waiting for myself, with a condenser and a water-air radiator by pipelines, fuel nozzles in the inlet pipes, connected by pipelines to a fuel pump, filter and flow regulator for supplying fuel to the cylinders, characterized in that it contains heat-insulated parts of the combustion chamber (cylinder head (covers), pistons and gas cylinders), a boiler combined with a thermally insulated exhaust manifold, as well as steam inlet valves with swirls in the thermally insulated cylinder head (s), connected by a pipe water with a boiler, to which water is supplied by a feed pump through a water flow regulator. 31. The device according to p. 30, characterized in that it contains thermally insulated (s) gas (s) cylinder (s), connected to the outlet (s) valve (s) of this (them) cylinder (s) with thermally insulated (and ) with the boiler steam line (s), and a condenser with heat-insulated (s) heat pipe (s) connected to the outlet valve (s) through the check valve (s) preventing the steam or gas from flowing back into the cylinder (s) ), while the outlet (s) valve (s) begins (s) to open at top dead center.

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