WO2018203766A1 - Direct and inverse stirling cycle having controlled heat transfer conditions in regenerator - Google Patents

Abstract

In hot purging of a Stirling cycle regenerator, a spent working medium is under constant pressure (having a high specific heat capacity), while in cold purging of a regenerator, when a working medium is heated, the working medium is in a constant volume state (having a low specific heat capacity). As a result, in the case of cold purging, a portion of the heat of the spent working medium inevitably is not required by the heated working medium from the regenerator, and said portion of the heat has be removed from the cycle to the heat engine cooler. The efficiency of the cycle drops lower than that to be expected in the case of the complete regeneration of heat in a regenerator. A technical solution is proposed which eliminates the deficiencies of a direct Stirling cycle, including a technical solution for improving the efficiency of an inverse Stirling cycle. As a result, the efficiency of the direct Stirling cycle is maximized, making it possible to use the present invention to create a multi-temperature heat-to-work converter, and to significantly increase cooling efficiency, in particular, in the case of heat engines using the inverse Stirling cycle.

Description

 Forward and reverse Stirling cycle with controlled heat transfer conditions in the regenerator.

 1. The technical field to which the invention relates.

 The invention relates to the field of “Steam engine”, since a gas acting as a working fluid of a heat engine is a vapor of a certain substance (liquid air, argon, neon, helium, etc.).

 2. The prior art.

 An analogue of the invention is the thermodynamic cycle of the Stirling engine, source of information: [J. Keenan, “Thermodynamics”, translated from English by A.F. Kotina edited by M.P. Vukalovich, State Energy Publishing House, Moscow-Leningrad, 1963, page 94, total 280].

 The thermodynamic cycle of the Stirling engine involves the regeneration of heat from the spent working fluid into a heated working fluid.

 The heat exchange is provided in a heat exchanger-heat regenerator, in which, when it is hot-blown by an exhausted working fluid, heat is accumulated, and during reverse cold purge by a heated working fluid, heat is transferred from the regenerator to the heated-working fluid.

Figure 1 shows the change in the volume V _r over the main piston, and the volume V _x over the displacing piston, i.e., the dynamics of the volume (V _r ) of the hot cylinder and the dynamics of the volume (V _x ) of the cold cylinder, which follows from the corresponding source of information: [http://rosavto42.ru/content/termodinamika/dvig stirling.html. Electronic resource, Article on the Internet, “Stirling Engine”, Fig. 10.23.].

 In FIG. 1 presents stage 1 2 of the thermodynamic process - this is the stage of compression of a cold working fluid and moving it through a heat exchanger-regenerator (hereinafter referred to as the regenerator) into a hot cylinder.

This is the so-called cold purge of the regenerator. The thermodynamic process proceeds with an increase in pressure under conditions of a decrease in volume above the displacing piston and some increase in volume above the working piston. The volume V _r does not change (almost), and the volume V _x decreases.

Under such conditions, the working fluid (gas) of the heat engine demonstrates the heat capacity C _v such that the following relations are satisfied:

Cy ^ Cp, Cy ⁼ Cp / g, where r is the adiabatic index of the gas.

At a low specific heat capacity, Su gas cannot take all the heat out of the regenerator at the same temperature difference, i.e. the very heat that in the regenerator, at the same temperature differences, was left by the exhaust gas during the hot purge of the regenerator. Step 2 3 shown in FIG. 1 is a working move. Isothermal expansion occurs and useful work is done.

The volume above the displacing piston (V _x ) does not change (almost), and the volume (V _r ) above the working piston increases.

 Stage 3 4 - at this stage, the displacing cylinder sucks the spent body, because its volume (Vx) increases. In addition, the pressure in an increasing volume drops, because the working fluid enters the cold cylinder, where the density of the working fluid increases.

Stage 3 4 is the stage of starting the hot purge of the regenerator, and at this stage the specific heat C _P acts.

Thus, the working fluid (gas) of the heat engine passes through the regenerator in an increasing volume mode with a pressure drop. Under such conditions, the gas exhibits a specific heat capacity С _Р and with the corresponding heat capacity of the materials of the regenerator, the gas can leave a significant amount of heat in the regenerator.

 Stage 4 1 - at this stage, the working fluid is squeezed out of the master cylinder through the regenerator - the hot purge continues in constant pressure mode, because the working fluid, although it falls into a small volume above the displacing piston, but fits there at a constant pressure, since during cooling the density of the working fluid increases, and the unit mass of the working fluid occupies a smaller volume.

 At this stage, the specific heat of the working fluid is also manifested in its value Cf.

Then follows the repetition of the thermodynamic cycle, starting from stage 1 2 and further according to the above algorithm.

 It turns out that, taking into account the understanding of the specific heat capacity of the working fluid, with hot and cold purges of the regenerator, the working fluid can take less heat (in the same temperature range) from the heat that is transferred to the regenerator of the heat engine.

The theoretical coefficient of regeneration (K _REG ) of heat (without taking into account the heat pressure in the heat transfer) cannot be more than the value:

 For monatomic gases, such as helium,

 HELIUM CREG = 1 / g = 1 / 1.66 = 0.6.

This means that 40% of the waste heat after a working thermodynamic cycle must be removed from the cycle, and then the efficiency of the Stirling engine (Efficiency _STIRLING ) is limited to the following limits:

 KPDSTIRLING ~ 60%.

In the actual thermodynamic processes, the effectiveness of the Stirling cycle will be even less, including due to energy losses in the regenerator due to the non-zero temperature pressure between the working fluids exchanging heat. 3. Disclosure of the invention.

 The technical problem solved by the present invention is to increase (increase) the efficiency of a heat engine that performs (implements) a direct Sterling cycle in engine mode, as well as increase (increase) the refrigeration coefficient of a heat engine that performs (implements) a reverse Sterling cycle in heat pump mode, by performing (implementing) a heat engine, respectively, a direct or reverse Stirling cycle with controlled heat transfer conditions in the regenerator.

 The technical result is achieved by the fact that in the claimed direct and reverse Sterling cycle with controlled heat transfer conditions in the regenerator, consisting of thermodynamic processes of heating and compressed working fluid, expansion and cooling of the working fluid, heat transfer from the spent working fluid to the heated working fluid, cooled working fluid and the heated working fluid is transferred to the regenerator of the heat engine under conditions when both working fluids, heated and cooled, are in a constant volume state.

 In addition, the technical result is achieved by the fact that in the claimed direct and reverse Sterling cycle with controlled heat transfer conditions in the regenerator, to compensate for heat losses in the regenerator of the heat engine, arising due to the non-zero value of the temperature head during the transfer of thermal power (during hot purging) heat capacity the material that accumulates heat in the regenerator, set the maximum possible, and with cold blowing of the regenerator, the heat capacity of the heat storage substance in the regenerator is set pour the lowest possible.

 The technical result of one of the options for implementing the Sterling cycle with controlled heat transfer conditions in the regenerator is also achieved by the fact that

 - at each of the four alternating stages of the cycle, two complementary thermodynamic processes are carried out simultaneously with the working fluid in the first and second devices for changing the volume of the working fluid — expanders, which are structural elements (working technical units) of the Sterling engine, namely:

 - at the first stage of the cycle, in the first expander, the working fluid is isochorically heated, and in the second expander, the working fluid is isochorically cooled;

 - at the second stage of the cycle, the working fluid is expanded in the first expander to obtain useful work, and the working fluid is compressed in the second expander to obtain useful work;

 - at the third stage of the cycle, the working fluid isochorically heated in the second expander, and the working fluid isochorically cooled in the first expander;

 - at the fourth stage of the cycle in the second expander expand the working fluid to obtain useful work, and in the first expander compress the working fluid and also to obtain useful work,

in addition, in isochoric processes of heating / cooling the working fluid of the first and third stages of the cycle, the working fluids are moved towards each other in the heat exchanger, preventing them mutual penetration, but at the same time create conditions so that heat flows from the working fluid previously heated in the previous cycle to the working fluid previously cooled in the previous cycle, provided that the working fluid previously heated in the previous cycle process and passed the heat exchanger is additionally cooled to a lower temperature in the cycle, and the working fluid, previously cooled in the previous process of the cycle, is additionally heated to the upper temperature in the cycle.

 The technical result is also achieved by the fact that in the direct and reverse Stirling cycle with controlled heat transfer conditions in the regenerator of the heat engine, when the claimed thermodynamic cycle is realized, the compression and expansion isoentropes are close to the isotherms, which significantly increases the efficiency of the traditional direct and reverse Stirling thermodynamic cycle, in which conditions for the transfer of heat are not controlled by the method proposed by the present invention.

 In gas regenerative machines, in order to perform (implement) heat pumps, cryogenic machines and moderate cold machines, the reverse Stirling cycle is used.

 The technical result of the implementation of the reverse Stirling cycle with controlled heat transfer conditions in the regenerator is achieved by the fact that, by analogy with the reverse Stirling cycle with the traditional heat regenerator, in the claimed Stirling cycle with controlled heat transfer conditions in the regenerator, the stages of the cycle (sequence of operations with the working fluid of the heat engine ) perform in the reverse order (in reverse order), realizing (performing), thereby, on the basis of the stated cycle, the heat pump.

 The cooling coefficient of such a heat pump that performs (implements) the declared reverse Stirling cycle with controlled heat transfer conditions in the regenerator will be several times larger than the cooling coefficient of a heat pump implementing the reverse Stirling cycle with a traditional regenerator installed between the working and displacing cylinder, which will be shown below.

 The claimed Stirling cycle with controlled heat transfer conditions in the regenerator, performed (sold) by a heat engine operating in engine mode, allows to reduce to zero the amount of waste heat that is traditionally transferred from the heat engine to the environment.

The specified technical result of a significant reduction in the amount of waste heat, which is traditionally transferred from the heat engine to the environment, is achieved by the fact that in a heat engine operating on a direct Stirling cycle with controlled heat transfer conditions in the regenerator, the working fluid after the end of the working stroke (at the cooling stage in a cycle), sent to a cold heat pump regenerator, which is paired with the temperatures of the cold and hot regenerator, respectively, with the temperatures of the waste heat and melting heat in the cycle, and the working fluid in the cycle, the heating step of the Stirling cycle heat transfer with controlled conditions in the regenerator, not heated only by supplying heat, but they are also heated in a hot regenerator of the aforementioned conjugate heat pump.

 The thermodynamic cycle can almost always be clearly illustrated by a graph that displays the parameters of the working substance (working fluid) at specific points of the cycle, and the transition paths (processes - steps) that connect these specific points of the thermodynamic cycle.

 If the parameters of the special points of the cycle are selected from the conditions of useful value of the cycle, then the schedule of the thermodynamic cycle helps in the disclosure of the invention and in its usefulness and significance.

 The claimed thermodynamic Stirling cycle with controlled heat transfer conditions in the regenerator allows you to efficiently convert low-temperature heat into mechanical work with an efficiency that is much higher than the efficiency of the Carnot cycle (efficiency) -

The claimed Stirling cycle with controlled heat transfer conditions in the regenerator, the graph of which is presented in FIG. 2, consists of several stages (thermodynamic processes) performed in the claimed cycle in the following order:

 1 2 - stage of expansion of the working fluid;

 2 3 - stage of isochoric cooling of the working fluid;

 3 4 - the stage of compression of the working fluid due to the suction of the piston into the cylinder volume, in which the working fluid was previously cooled (thermal compression);

 4 1 - stage of isochoric heating.

 The thermodynamic cycle provides for the regeneration of heat from the spent working fluid into a heated working fluid.

 The Stirling cycle is known, where the regenerator stores heat during hot purging in its heat capacity. During cold blowing, the working fluid takes away part of the heat from the regenerator, and the rest of the heat from the regenerator must be emitted, transferred to the environment.

 It is also known that in the Stirling engine, the structural elements (working technical units) are expanders - two cylinders and two pistons - working and displacing, and the volumes above the pistons are connected through a regenerator.

 According to the Stirling cycle with controlled heat transfer conditions in the regenerator, it is possible either to switch the regenerator from the purge in step 2 of the 3 cycle to the purge in the opposite direction in step 4 of the 1 cycle, as shown in FIG. 2, but it’s simpler, without switching the regenerator, to transfer the heat of the spent fluid from stage 2 3 of the cycle to the stage of heating of the working fluid 4 of 1 cycle, but in a different cylinder volume, where the claimed Stirling cycle is also carried out with controlled heat transfer conditions in the regenerator, but with a shift working procedures of the thermodynamic cycle in time.

 When step 2 3 is performed in the first cycle, step 4 1 is performed in the second cycle.

When the work in step 1 2 is performed in the first cycle, then the work in step 3 4 is performed in the second cycle. If a different volume used for the implementation of the cycle is formed above the reverse side of the double-acting piston, then in the first volume, in stage 1-2 of the cycle, for example, the next stage of the cycle is performed, which produces useful mechanical work, and in the second volume, on the other side of the piston, stage 3-4 is performed. Formally, this is the compression stage, and in the mathematical evaluation of the work performed, a sign is obtained for this work that is opposite to the work sign in stage 1 2 of the cycle, but in fact, the previously discharged working fluid is compressed, in which pressure reduced from the average level is obtained by isochoric cooling. In this case, at the stage of the cycle of such compression of the working fluid, the piston is drawn into this second volume, and the pulling force works against external forces, increasing useful work.

 The isochoric cooling step 2 3 cycles and the isochoric heating 4 1 cycle shown in FIG. 2, is provided by the transfer of heat from step 2 of the 3 cycle to step 4 of the 1 cycle using a heat regenerator (heat exchanger for exchanging heat with the working fluid of another thermodynamic cycle). The regenerator at the stage 2 3 of the cycle (hot purge) takes the heat, and at the stage 4 of the 1 cycle (cold purge) gives the heat to the heated working fluid.

3.1. Air parameters at specific points of the Stirling cycle with controllable

 heat transfer conditions in the regenerator.

A graph of the Stirling thermodynamic cycle with controlled heat transfer conditions in the regenerator is shown in FIG. 2., where the temperature of the supply of heat supply T _GO p = Tj = 340 ° K, and the temperature of the removal of waste heat T _XO l = T ₃ = 300 ° K.

The temperature Thol ⁼ 300 ° K is selected above the typical value of the temperature of the natural reservoir for waste heat from the operation of heat engines.

The temperature T _GO p = 340 ° K was selected on the basis of the waste heat parameters from combined cycle plants (CCGT) for reasons of expediency of converting low-grade waste heat of technological processes, which, for example, CCGTs are emitted into cooling towers or in reservoirs, into useful mechanical work. At a low temperature of the coolant (70 ° C) circulating in the secondary cooling circuit of the CCGT steam turbine condenser, such a coolant is not used for heating buildings and structures, and the heat from such a coolant is transferred to the environment by means of cooling towers or reservoirs with a large area of the reservoir mirror.

 It will be shown below that the CCGT waste heat can be converted with a sufficiently high coefficient into useful mechanical work.

In these calculations for work and heat at the stages of the declared thermodynamic cycle, in the temperature range of the heat engine under discussion (T _G OR + T _X ol) _> air parameters are presented as the working fluid of the heat engine, which follow from the information in the tables presented in the information source [1]: ["Reference for thermophysical properties of gases and liquids ”, N. Vargaftik, Nauka, Fizmatlit, Moscow, 1972,“ Air ”, pp. 596-597].

In this regard, in accordance with the numerical values of V - volume, and T - temperature, fixed for points 2 and 4 of the cycle in the graph presented in Fig. 2, for points 2 and 4 of the cycle from the source of information [1] the numerical values of the entropy S ₂ and S ₄ for air, namely:

- for point 2 of the cycle - volume V2 = 5.167 l / kg, temperature T ₂ = 340 ° K

corresponds to the value of entropy S ₂ = 5.392 kJ / kg * deg;

- for point 4 of the cycle - volume V4 = 4.446 l / kg, temperature T ₄ = 300 ° K

corresponds to the value of entropy S ₄ = 5.238 kJ / kg * deg.

 In the initial state of the heat engine, we have the following initial conditions:

the pressure above the piston and under the piston is equal to the initial pressure P ₀ , the volume matters - V ₀ , the temperature matters - T ₀ , namely:

- pressure P ₀ = 20.0 MPa = 200.0 bar,

 - temperature To = 320 ° K,

- volume V ₀ = 4.810 l kg.

 The pressure (Pi) at point 1 of the cycle is determined according to the law of Kliperon-Mendeleev:

Pi = P ₀ * Ti / T ₀ = 20.0 * 340/320 = 21.25 MPa = 212.5 bar.

 The pressure (Pg) at point 3 of the cycle is determined according to the law of Klaiperon-Mendeleev:

Pb = Po * T ₃ / T ₀ = 20.0 * 300/320 = 18.75 MPa = 187.5 bar,

where Ti = 340 ° and T ₃ = 300 ° K is the temperature, respectively, at points 1 and 3 of the cycle.

 The entropy parameters of points 1 and 3 of the cycle are determined by interpolation from the values of entropy obtained from the information source [1], from the corresponding pressure and temperature values closest to points 1 and 3 of the cycle, namely:

- for point 1 of the cycle correspond: temperature T340 = 340 ° K and pressure Pi = 21.25 MPa, while the closest section “temperature-pressure” at temperature T340 = 340 ° K and at pressure P ₃₀ about = 300.0 bar = 30.0 MPa corresponds to entropies for air S ₃₀ o 340, namely:

 S3oo_34o = 5.247 kJ / kg * deg;

- for point 3 of the cycle correspond: temperature T300 = 300 ° K and pressure Pz = 18.75 MPa, while the closest cross-section “temperature-pressure” at temperature T300 = 300 ° K and pressure P = 150.0 bar = 15.0 MPa corresponds to the value of entropy for air Si ₅ o 300, namely:

 Si5o_3oo = 5.343 kJ / kg * deg.

 Explanation: in the tables of the information source [1] there are no “temperature-pressure” sections for pressure and temperature for air parameters that are closer to the parameters of the declared cycle.

 Based on this, we have the following results:

- after interpolation for point 1 of the cycle we have the following values: at pressure Pi = 21.25 MPa and temperature Tz ₄₀ = 340 ° K, the result of entropy (Si = S21 _. 25J40) should be close to the value of entropy S ₂ = 5.392 kJ / kg * deg, (taking into account that Pzoo ⁼ 300.0 bar at Tz4o ⁼ 340 ° K) [1], namely:

S21.25 _. 340 = S ₂ - (S ₂ - S ₃₀₀ j4o) * 1 · 25 / (Rzoo - Po) = 5.392 - (5.392 - 5.247) * 1.25 / (300.0 - 200.0) =

 = 5.392 - 0.145 * 1.25 / 100 = 5.392 - 0.002 = 5.390 kJ / kg * deg,

Si = S ₂ i 25j4o = 5.390 kJ / kg * deg;

 - after interpolation for point 3 of the cycle we have the following values:

at a pressure of Pz = 18.75 MPa and a temperature of T300 ⁼ 300 ° K, the result of entropy (S ₃ = S18.75J00) should be close to the value of entropy S4 = 5.238 kJ / kg * deg, (given that P = 150.0 bar at T300 = 300 °) [1], namely:

Si8 _. 75_3oo = S ₄ + (S, 5o_3oo - S ₄ ) * 1.25 / (Po - Piso) = 5.238 + (5.343 - 5.238) * 1.25 / (200.0 - 150.0) =

 = 5.238 + 0.105 * 1.25 / 50 = 5.238 + 0.002 = 5.240 kJ / kg * deg.

S3 = Si ₈ .75_3oo = 5.240 kJ / kg * deg.

Entropy for air at points 1, 2, 3, 4 of the cycle, respectively, has the following meanings:

Si = 5.390 kJ / kg * deg; S ₂ = 5.392 kJ / kg * deg; S ₃ = 5.240 kJ / kg * deg; S ₄ = 5.238 kJ / kg * deg.

3.2. Calculation of the stages of the claimed Stirling cycle with controlled heat transfer conditions in the regenerator.

In the graph of FIG. 2, at the working stage of stage 1 2, mechanical work Li _-2 is generated from thermal energy.

 To calculate the work, you need to know the change in the volume of the working fluid and the change in pressure at the stage of the thermodynamic cycle.

The coefficient of change in pressure during expansion (K _R dssh) at stage 1 2 of the cycle and the coefficient of change in pressure during compression - suction (CSJAT) at stage 3 of 4 cycles are slightly different from each other, namely:

 EXT = Pi / Po = 21.25 / 20 = 1.062

 XJAT = Po / Rz = 20 / 18.75 = 1.066

 The volume Vi at point 1 of the cycle is calculated as:

Vi = V ₂ / KRASH = 5.167 / 1.062 = 4.865 l / kg.

 The volume V3 at point 3 of the cycle is calculated as:

V ₃ = V ₄ * XJAT = 4.446 * 1.066 = 4.739 l / kg.

The work of expansion (Lp _A cin) of the working fluid with pressure from P] to pressure P ₀ with a coefficient of expansion with pressure K _RA ssh ⁼ 1,062 is calculated according to the formulas from the information source: [M.P. Vukalovich, I.I. Novikov, “ Technical thermodynamics ”, Energy, Moscow, 1968], namely:

 V2 5.167

LpAcm = J ((const, /) - P) dX = J ((4.865 * 21.25 / X) - 20.0) dX = 0.1835 ~ 0.184 kJ / kg, v, 4.865 where: const] = V ₁ * P ₁ ; P ₀ = 20.0 MPa; Pi = 21.25 MPa — initial expansion pressure; Vi = 4.865 l / kg - initial volume of expansion.

The piston suction work (L _B CAC) according to the pressure from P3 to the pressure P0 into the volume previously discharged in the coefficient of CSF ⁼ 1,066 times, is calculated as:

 4 4.446

LBCAC = J ((const- / LH) - Po) dX = j ((4.739 * 18.75 IX) -2ΰ.ϋ) άΧ = 0.1892 - 0.189 kJ / kg, ν ₃ 4.739

where: const ₂ = V ₃ * P ₃ ; P ₀ = 20.0 MPa; P ₃ = 18.75 MPa; V ₃ = 4.739 l / kg.

Taking into account the fact that suction and expansion creates forces acting in one direction (on a double-acting piston), the full work of the piston stroke L _X ODE ⁼ L is equal to the sum:

ЦИ CYCLE ⁼ I LpAciu I + I LBCAC \,

where the full work of the piston stroke is equal to the sum of the work of the stages of the cycle, taken modulo:

 L cycle = I 0.184 I + I 0.189 I = 0.373 kJ / kg.

 In regenerative cycles, when they are calculated, they necessarily examine the heat that the thermodynamic process is able to transfer, and they examine the situation to see if all the transferred heat has a temperature higher than the temperature of the working fluid at the heating stage.

Without taking into account the losses, the heat at stages 4 1 and 2 3 of the cycle is supplied in the amount of CYCLE ⁼ Q4 1 + + h _j ⁼ 0.373 kJ / kg, and the energy from the cycle at stages 1 2 and 3 4 is removed in the form of mechanical work in the amount of + = 0.373 kJ / kg., I.e. > CYCLE ⁼ CYCLE-

3.3. Calculation of the heat of the stages of the Stirling cycle with controlled conditions

 heat transfer in the regenerator.

 From the graph shown in FIG. 2, it follows that the heat (Q) of each stage of the declared thermodynamic cycle can be estimated through the entropy of points 1, 2, 3, 4 of the cycle, given that the adiabatic index for air is r = 1.41.

The heat in step 4 1 (Q _{4 j} ) is calculated as:

Q _4J = 0.5 * (Si - S ₄ ) * (Ti + T ₄ ) / g = 0.5 * (S _{21 25 340} - S ₄ ) * (T, + T ₄ ) / g = 0.5 * (5.390 - 5.238 ) * (340 + 300) / 1.41 = 0.5 * 0.152 * 640 /1.41 = 34.496 kJ / kg.

The heat in step 2 3 (Q ₂ 3) is calculated as:

Q ₂ s = 0.5 * (S2 - S3) * (T ₂ + S ₃ ) / g = 0.5 * (S ₂ - S „.75_30o) * (T ₂ + T ₃ ) / g = 0.5 * (5.392 - 5.24) * (340 + 300) / 1.41 = 0.5 * 0.152 * 640 / 1.41 = 34.496 kJ / kg.

 Conclusion: the heat capacity of stage 4 1 and stage 2 3 completely coincided.

Q ₄ 1 = Q _{2 j} = 34.496 kJ / kg.

 Without taking into account losses in the regenerator, there is no waste heat at the stage 2 3 of the cycle - it is completely absorbed by the heating stage 4 1.

 Only useful work is derived from the claimed thermodynamic cycle

3.4. Evaluation of the efficiency of the direct and reverse Stirling cycle with controlled heat transfer conditions in the regenerator. It is known that heat engines operating on the forward and reverse Sterling cycles are used, respectively, both in the design of engines and in the design of heat pumps.

 To evaluate the effectiveness of the claimed direct and reverse Sterling cycles for the same temperature ranges of heat supply and removal in the direct and reverse cycles, it is possible by quantitative comparison of the corresponding values of the prototype cycle parameters - a heat engine made (working) on the direct and reverse Carnot cycle.

3.4.1. Evaluation of the efficiency and effectiveness of the claimed real direct Stirling cycle with controlled heat transfer conditions in the regenerator of a heat engine made (working) in engine mode.

 The efficiency of the thermodynamic cycle of a heat engine operating in engine mode is determined as the ratio of useful mechanical work to all the energy output (C> output) from the thermodynamic cycle:

QjBblBOfl ⁼ bcyclde + QEPOC

where Ovod is the energy derived from the cycle, calculated as the sum of the useful mechanical work (CYCLE) and the waste heat (Q _B poc)

 The above calculation of the cycle showed that in the absence of losses in the regenerator, from the Stirling cycle with controlled heat transfer conditions in the regenerator, waste heat is not removed, since all of it (heat) is transferred inside the cycle to the heated working fluid. Energy is removed from the cycle only in the form of mechanical work, i.e. there is no waste heat in the cycle, and only useful work remains in the cycle.

 Thus, the theoretical cycle efficiency (KPTTEORET), In percent, is equal to:

 EFFICIENCY = YuO * ЦИ CYCLE / OVSHOD = 100 * 0.373 / 0.373 = 100%.

 Efficiency = 10%.

 Under real conditions, heat should be removed from the cycle, which is not transferred to the heated working fluid from the spent working fluid in the regenerator, and this heat is all the more, the greater the temperature (thermal) pressure required for the used regenerator in order to transfer the amount of heat (OREGEN) FOR the above example of loop calculation:

QpErEH = Q ₄ _i = Q ₂ j = 34.496 kJ / kg.

The regenerator must have a thermal power Q _PEEEH ⁼ 34.496 kJ per kilogram of circulating working fluid.

If the volume of the working cylinder decreases from the initial volume Vo = 4.81 liters, for example, by half, the thermal power of the selected regenerator can also be reduced by half, but if the regenerator is left the same, then the temperature incursion in the regenerator will also be reduced by half. Reducing the temperature rise in pressure in the regenerator leads to an increase in the efficiency of the thermodynamic cycle due to the reduction of losses in the regenerator.

 The efficiency of the declared real cycle (efficiency at) of Stirling with the real temperature incursion (dT) during controlled heat transfer in the regenerator according to the declared cycle is determined by the following ratio:

Efficiency JT = ЦИ CYCLE / (ЦИ CYCLE + Q <t_i * dT / (T _GO P - Thol)),

where dT is the temperature (thermal) pressure in the regenerator when transmitting the necessary heat power.

For example, for the temperature head dT = 0.5 degrees, the value for the efficiency (

The cycle is the following value:

Efficiency _ (} m = 0 ₅ = 0.373 / (0.373 + 34.496 * 0.5 / 40) = 0.373 / (0.373 + 0.431) = 0.373 / 0.804 = 0.4639, where: Q ₄ , = 34.496 kJ; (T _Г0Р - Т _Hall ) = (340 ° K - 300 ° K) = 40 ° K.

 The efficiency of the declared real direct Stirling cycle is calculated similarly for other values of the thermal head (dT), and these efficiency values are presented in Table 1., column 2., namely:

Efficiency _DT = 1 = 30.20% and Efficiency _DT = 2 = 17.79%

The efficiency of the prototype

direct cycle in the considered temperature range (T _GO p ^ T _X ol) is estimated by the Carnot formula:

KPDKARNO ⁼ 100 * (Tgor - cold) / T _GO p

SCHROTOTYPE = 100 * (T _GO p - T _cold ) / T _GO p = 100 * (340-300) / 340 = 11.76%

Efficiency_G1ROTOTYPE ⁼ 11.76%

 Table 1., column 3, presents the values of the efficiency of the direct cycle of the prototype, the same for all three rows of Table 1.

Taking into account the above estimates of the efficiency (efficiency_at) of the declared real direct Sterling cycle, and the calculation of the efficiency of the direct prototype cycle (KPD_PYUTOTYPE _A ) _{> it is} possible to quantify the efficiency (EFF _motor (1t) of the declared real direct Sterling cycle, a heat engine running in engine mode, for a specific temperature head (dT) in the regenerator in comparison with the efficiency of the prototype (EFFICIENCY) ₅ determined by the following ratio:

For example, for the heat pressure dT = 0.5 degrees, according to this ratio, the efficiency of the declared real direct Stirling cycle, as shown above, is efficiency_t = 0.5 ⁼ 0.4639, whence the efficiency value (EFF_wig_at_at _{= 0} 5) of the declared direct cycle with thermal pressure dT = 0.5 is defined as:

0.5 / KPT_P TOTYPE ⁼ 0.4639 / 0.1 176 = 3.945 times. 3.4.2. Evaluation of the refrigeration coefficient and efficiency of the claimed real reverse Stirling cycle with controlled heat transfer conditions in the regenerator of a heat engine made (working) in the heat pump mode.

Based on the conditions of reversibility of thermodynamic processes, it should be assumed that the efficiency (EFF_T _N T) of the reverse Stirling cycle with controlled heat transfer conditions in the heat generator regenerator made (operating) in the heat pump mode will be equivalent to the efficiency (EFF_ ^ ig_at) of the direct Stirling cycle with controlled heat transfer conditions in the regenerator of the heat engine, made (working) in the engine mode, namely:

 EFF t n dT = EFF _motor _ () t

 When dT = 0.5, the efficiency of the claimed Stirling cycle in the heat pump mode has the following value:

EFF_ _{t n} _ <yy = o.5 = EFF _jjBHT_dT = 0 5 = 3.945 times.

 Accordingly, for the thermal pressure dT = 1, the efficiency of the claimed Stirling cycle in the heat pump mode is:

 EFF T H_dT = i = 2.568 times,

and for the thermal pressure dT = 2, the efficiency of the claimed reverse Stirling cycle is:

EFF _T H_dT = 2 ^~ 1.13 times.

The refrigeration coefficient (K _X OL_PROTOTYPE) of a heat pump operating in the reverse Carnot cycle (prototype) is calculated by the inverse Carnot formula as follows:

T _XO l / (TGOR - Thol) = 300 / (340 - 300) = 7.5.

Table 1., column 4, presents the values of the refrigeration coefficient of the reverse cycle of the prototype (CHOL _^ PROTOTYPE), the same for all three rows of Table 1.

The refrigeration coefficient (Khol_t _. N at) of a heat pump made according to the declared real reverse Stirling cycle with controlled heat transfer conditions in the regenerator with the corresponding heat head (dT) is determined by the following relation:

Khol_T H_dT ⁼ Prototype Prototype * EFF_t.N_ <YY

 For the thermal head dT = 0.5, we have the following value of the refrigeration coefficient

(Khol T H_dT = 0.5)

chill_tn <1T = o.5 = Khol _^ PROTOTYPE * EFF_tn_ (1T = o.5 = 7.5 * 3.945 = 29.59 times. The cooling coefficients of the declared real reverse Stirling cycle for thermal heads dT = 1 and dT = 2 are calculated similarly , respectively:

Kxcm_T.H_dT = 1 ⁼ CHOL_PUTOTYPE * EFF_t.n_at = ι = 7.5 * 2.568 = 1 .26

K _H ol T.N dT = 2 ⁼ Khol_PROTOTYPE * EFF t H_dT = 2 ⁼ 7.5 * 1.5 13 = 1 1 .35

Table 1., Column 5, presents the values of the refrigeration coefficient (K _X ol of the declared cycle in the heat pump mode at different values of the thermal head (dT). Table 1. presents the parameters of thermal devices implemented on the basis of the direct and reverse Stirling cycle with controlled heat transfer conditions in the regenerator at the highest temperature in the declared cycle (T _G RR ⁼ 340 ° K), and at the lowest temperature in the cycle (Thol ⁼ 300 ° K) at the corresponding temperature head (dT) in the regenerator.

 Table 1.

On the basis of a heat engine and a heat pump, made according to paragraph 6 of the method of the claims, it is possible to create a mono-temperature converter of heat to work.

 But the addition of a heat pump to a heat engine increases the cost of such a heat converter in the work.

 The greater the output of a heat engine with a natural refrigerator (without a heat pump) in relation to the amount of work spent on the heat pump drive (Mt n), the higher the cycle efficiency.

 The higher the efficiency of the cycle, the lower the cost of equipment when a heat pump is introduced into it (into the equipment design).

Increased cost factor (Kudrrozh_ (1t) is proportional to the total work cycle and the heat drive pump to work cycle with natural refrigerator. In evaluating the heat pump of the drive on the reverse formula Carnot operation of the heat pump drive (M _{t n)} entraining the entire junk heat from - under the cycle to the heater, it is determined by the estimate (with an overestimation of the necessary work) as a percentage of the heat input to the cycle, as follows:

M _so- called ^ YuO - KSh t / Khol PROTOTYPE

Then the cost of equipment (K _UD orozh_at) can be determined as:

Kudr m = (M _{t n.} + Efficiency atm) / _d T = efficiency

= ((100 - efficiency _dT) / K _cold PROTOTYPE) Efficiency + m) / efficiency _dT. For a cycle with thermal pressure dT = 0.5 degrees and efficiency at = 0.5 ⁼ 46.39%, rise rate

 ((100 - 46.39) /7.5 + 46.39) / 46.39 = (53.61 / 7.5 + 46.39) / 46.39 =

 = (7.15 + 46.39) / 46.39 = 53.54 / 46.39 = 1.15.

 For a cycle with a thermal pressure dT = 1.0 degrees and efficiency_ (П- = 1.о = 30.20%, the appreciation coefficient according to the formula is:

(69.8 / 7.5 + 30.20) / 30.20 = (9.31 + 30.20) / 30.20 = 39.51 / 30.20 = 1.31. For a cycle with thermal pressure dT = 2.0 degrees and efficiency at = 2 0 = 17.79%, the cost factor is:

82.21 / 7.5 + 17.79) / 17.79 = (10.96 + 17.79) / 17.79 = 28.75 / 17.79 = 1.62. Table 1, column 6, presents the values of the coefficients of appreciation

mono-temperature converter of heat into mechanical work (without taking into account the savings usually allocated for the construction of a cooling tower).

Since the rise in price coefficients (K _U d ₀ ry__t) do not approach infinity, it should be recognized that all variants of the claimed Stirling cycle with controlled heat transfer conditions in the regenerator, presented in Table 1., are guaranteed to be realized.

 The feasibility of monotemperature converters is facilitated by the claimed Stirling cycle with controlled heat transfer conditions in the regenerator, which provides a high value of efficiency with a small temperature difference in the heat supply and removal in the cycle.

 With such small temperature differences in the cold and hot heat pump regenerator, the efficiency of the heat pump itself will be very high.

 As a heat pump, a heat engine operating according to the reverse Stirling cycle, the shaft of which is rotated by an external force in the direction opposite to its rotation in a heat engine made in the engine mode, is increasingly being used today. A heat pump made according to the reverse Sterling cycle does not depend on the condensation temperature of freon and generally does not require freons and freon-like working fluids, designed strictly for certain temperatures. Source of information: [Tanklevsky V.I., Gruzman P.M., Kirillov N.G., Sudar Yu.M. “Decentralized heat supply systems with heat pumps operating on the reverse Stirling cycle”, Heat-energy-efficient technologies, Information Bulletin, N 1, S-Pb., 1997, p. 38-40].

 On the other hand, the efficiency of such a heat pump itself can also be significantly increased if, in a heat engine (heat pump) operating on the reverse Stirling cycle, the declared reverse Stirling cycle with controlled heat transfer conditions in the regenerator is applied (used).

The use of the claimed reverse Stirling cycle with controlled heat transfer in a regenerator for a heat pump of mono-temperature heat converters to work will reduce the cost increase of mono-temperature converters (in relation to equipment using a natural reservoir for discharging heat from the spent working fluid of the heat converter to work).

 The conclusion about the possibility of constructing mono-temperature converters using the inverse Sterling cycle with controlled heat transfer conditions in the regenerator does not contradict the laws of physics and the second law of thermodynamics in its modern probabilistic formulation given in the information source [2]: [M.P. Vukolovich, I.I. Novikov, “Technical Thermodynamics”, Energy, Moscow, 1968, Textbook for higher educational institutions].

 Page 97, [2]: “The second law of thermodynamics, according to modern concepts, is not an exact law of nature, similar to the laws of conservation of momentum or conservation of energy. The second law of thermodynamics has, as will be shown in detail in § 3–9, a statistical character and therefore it is only “average”.

 There is a difference in the thermodynamic and statistical formulations of the second law of thermodynamics, which consists in the fact that the statistical statement of the second law of thermodynamics claims that in a closed system processes accompanied by an increase in entropy are the most probable, while the thermodynamic statement of the second law of thermodynamics considers such processes to be the only possible ones. .

 This difference is very significant: the statistical formulation of the second law of thermodynamics not only does not deny, but, on the contrary, suggests the possibility of processes as a result of which the system passes from more probable states to less probable, with a decrease in entropy, while the thermodynamic formulation of the second law of thermodynamics completely excludes the possibility similar processes.

 Page 98, [2]: "In nature, along with the dissipation of energy, reverse processes always occur, as a result of which new types of energy arise from" scattered "energy, for example, the energy of electric charges, the energy of vortices and tornadoes, the energy of excitation and decay of atoms, etc." .

 Page 56, [2]: "Based on the law of conservation of energy, it is acceptable to assume that any conceivable process that does not contradict the law of conservation of energy is fundamentally possible and could take place in nature."

 4. A brief description of the drawings.

4.1. Figure 1 presents graphs of the dynamics of changes in volume V _r (hot cylinder) and changes in volume V _x (cold cylinder) in a Stirling engine.

 4.2. In FIG. 2 presents one of the possible graphs of the thermodynamic cycle of a heat engine, made taking into account the claims, according to paragraph 1.

4.3. In FIG. 3. A graphical illustration is presented of an engine device with two oblique washers satisfying the method of claim 1. 4.4. In FIG. 4 presents the dynamics of changes in the working volumes of a cylinder with a double-acting piston in a heat to work converter arranged within the framework of the requirements of paragraph 1 of the method of the claims.

 4.5. In FIG. 5 shows a pneumatic block diagram of a device according to paragraph 1 of the method of the claims.

 5. The implementation of the invention.

 5.1. The implementation of the device according to the method of paragraph 2 of the claims.

 It is known that the magnetocaloric effect (FEM) and the heat capacity of magnets in the temperature range (298 ° K to 353 °) and in magnetic fields with magnetic field induction in the range (0.0 T 1.0 T) are experimentally investigated.

 In the corresponding works on the heat capacity of magnets, it was first established that for nanoscale stabilized magnetite in magnetic fluids, the magnitude of the magnetocaloric effect exceeds the magnitude of the FEM of magnetite in the microheterogeneous state:

 • nanoscale magnetite in magnetic fluids in the temperature range (336 ° K + 340 ° K) undergoes an order-to-order magnetic phase transition;

 • the heat capacity of nanoscale magnetite in magnetic fluids in a zero field exceeds the heat capacity of magnetite in a microheterogeneous state;

 • unstabilized fine magnetite passes into hematite during the oxidation process;

 • the heat capacity of a magnet strongly depends on the magnitude of the magnetic field.

 Source of information: [Arefyev Igor Mikhailovich, “Magnetocaloric effect and heat capacity of highly dispersed magnets”, 02.00.04 - Physical Chemistry, Abstract of the dissertation].

 In accordance with the principle of the magnetocaloric effect, the magnetic field in the material of the regenerator (in the material that stores heat) should be reduced to zero when the regenerator is hot-blown and the magnetic field in the regenerator material should be increased during cold-blown. In the regime of cold purging, the temperature of the material of the regenerator will increase, which will allow transferring all the heat stored by the regenerator to the working fluid of the Stirling engine, despite the fact that the specific heat of the working fluid in the regime of cold purging is reduced.

 5.1.2. The implementation of another variant of the device according to the method of paragraph 2 of the formula

 inventions.

In household refrigerators, the property of changing the heat capacity of the working agent with a change in pressure is used. At low pressure, the liquid agent evaporates at the temperature of the freezer and exhibits an increased specific heat, and at increased pressure the gas goes into the liquid phase, condensation heat is released at an increased temperature in the cooling radiator located on the rear wall of the refrigerator. If the evaporation of the regenerator agent is carried out during the hot purge of the regenerator, and the condensation of the regenerator substance is carried out at the time of the cold purge of the regenerator, then the heated working fluid of the Sterling engine will absorb all the heat previously accumulated by the regenerator substance.

5.2. Implementations of the device in the direct Stirling cycle mode according to the method of paragraph 1 of the claims.

 The easiest way to show the operability of the claimed thermodynamic cycle according to paragraph 1 of the claims on the example of an engine with a slanting washer.

 The engine with a slanting washer is a piston machine without a crankshaft. Most often, in such a machine, the piston, cylinder and piston rod are cylindrical and coaxial, and the rod has the ability to move along itself in the guides of the heads, if the piston is double-acting. In addition, the output shaft is located outside the cylinder, in the immediate vicinity of the cylinder. The output shaft has the ability to rotate around its axis in angular contact bearings, a washer is fixed to the shaft, and some part of the plane of the washer is not perpendicular to the axis of the shaft. In this case, the piston rod abuts against the washer and during the stroke of the piston with the rod, the rod affects the oblique portion of the washer and the washer together with the shaft rotates around the axis of the output shaft.

 A feature of the engine is the fact that the mass of the working fluid in the cylinder is always enclosed in the volume between the piston and the corresponding cylinder head. After the working stroke, the temperature of the working fluid remains high (too little internal energy is converted into mechanical work). And, despite the fact that the working fluid is considered to be spent in one part of the cylinder, the spent working fluid will always be hotter than the working fluid in another part of the cylinder, and if the heat from the spent is transferred to this, another part of the cylinder, and this gas volume is slightly cooled. , then there will be forces that will return the piston to its original state and at the same time will work against external forces.

 The appearance of some components of such an engine, designed to implement the claimed thermodynamic cycle, is shown in FIG. 3., which is a graphic illustration of the engine with two oblique washers.

 Pistons 1 and 1a are located on rods 2 and 2a. On the Za-3 output axis, two washers 4 lev and 4 right are placed, and between them the rods and pistons are located so that the ends of the rods 2 and 2a touch the surfaces of the two washers 4 lev and 4 rights on tracks 5 and 6, the plane of tracks 6 is perpendicular to the axis of the shaft Za-3 . The relative position of the rod 2a with the piston 1a and the washers, as shown in FIG. 3, such that the ends of the rod abut against that part of the profile of the surfaces of the washers, 5 that has an inclined plane with respect to the axis of rotation of the output shaft ZA-3.

When the output shaft Za-3 rotates in the direction ω (with this arrangement of rods and washers), the position of the piston 1 and rod 2 does not change (the piston with the rod cannot move along the axis of the rod), and the rod 2a with the piston 1a is forced to move in the direction of force F. The direction of the application of force F is opposite on the pistons and, since the rod also acts on the oblique surfaces of the opposite curvature of the two washers, the forces arising on the piston during the working stroke always rotate the shaft in one direction.

 When the pistons move, the working volumes between the piston and the cylinder heads change (the heads are not shown in Fig. 3, but the guides in the heads and rods are aligned and the heads are placed between the piston and the oblique washers).

 In FIG. Figure 4 shows the dynamics of changes in the working volume of the cylinders: between the piston and the left cylinder head and between the piston and the right cylinder head (when the output shaft rotates with an angular speed of ω / s).

 On the upper group of graphs is the change in the volume of the cylinder left and right of the piston 1.

The lower group of graphs shows the change in the volume of the cylinder left and right of the piston 1a.

With continuous rotation of the output shaft with washers of the described profile, the volumes under the left and right cylinder heads will change.

 And, on the contrary, with proper alternation of heating and cooling of the working fluid in the agreed volumes, forces will arise that cause the motor shaft to rotate continuously.

 The pneumatic structural diagram of a device operating according to the claimed cycle is shown in FIG. 5.

 The piston engine is shown in FIG. 5 by cylinder 7, piston 1, rod 2 and cylinder heads 8 left and 8 right with guides for rod 2. In FIG. 5. the position of the stem and the slanting washers of 4 levs and 4 rights is displayed, in which the ends of the stem 2 abut against the non-inclined part of the washers of 4 levs and 4 rights, i.e. in surface 6 (see FIG. 3 for more details).

 The piston may be hollow with an internal baffle. The piston cavity to the right of the septum, the cylinder volume Vnp and the right stem cavity are communicating vessels. The right cavity of the rod, in addition, communicates with external pipelines through connector x. Similarly, the piston cavity to the left of the septum, the cylinder volume Ulev and the left stem cavity are communicating vessels. The left cavity of the rod, in addition, communicates with external pipelines through the connector u.

 In the cylinder heads there are 8 levs and 8 rights, windows and corresponding connectors are provided for connecting external pipelines.

 Along the stem, in volumes, under the left 8left, and under the right 8leight, the cylinder head can move the partition - the curtain, respectively 9left and 9right. In the block diagram of FIG. 5, the drives for moving these partitions along the stem are not reflected (the technical solutions of the drive are infinitely many), however, the drives move the partitions from the piston to the head after making a stroke in this part of the cylinder and in the phase of piston suction from the head to the piston.

When moving the partitions, there is no significant pressure difference between the working fluid of the heat engine on the sides of the partitions, since the design of the device according to the circuit of FIG. 5 provides for free flow of the working fluid from the volume in front of the partition (from the volume, towards which the partition moves) into the vacant volume behind the partition. The working fluid flows through the corresponding cavities of the heat exchangers 10 (heat exchanger-regenerator), 11 (pre-heating heat exchanger) and 12 (pre-cooling heat exchanger) in a constant volume mode, but with a synchronously changing pressure before and after the partition.

 In FIG. 5, the state and position of the piston is fixed, which was preceded by the heating of the working fluid in the right part of the cylinder and the cooling of the working fluid in the left part of the cylinder. Moreover, even after the end of the working stroke, the pressure to the left and right of the piston became the same, but even when the piston shifted to the left from the middle position in the cylinder after the working stroke, the temperature of the working fluid in the left side of the cylinder is lower than the temperature of the working fluid in the right side of the cylinder.

 To organize the next working stroke, it is necessary that the working fluid on the left side of the cylinder is heated and there the pressure increases, and the temperature on the right side of the cylinder drops and there the pressure decreases. The heating and cooling procedures occur when the piston is locked in the extreme position (when the ends of the rods are not on the inclined surfaces of the oblique washer).

 So, the piston and the rod are fixed, and the curtain-partitions 9left and 9th right move from right to left, in the direction of arrows Pr1 “purge 1”.

 When blowing Pr1, in the right part of the cylinder with a heated working fluid, the gas moves from the connector J, through the check valve 13 to the nozzle d of the regenerator 10, then through the nozzle r to the fitting p of the heat exchanger 12. From the outlet of the heat exchanger 12 (fitting c) enters the vacant volume behind the curtain 9 rights, but enters already at the lowest temperature in the cycle.

 When blowing Pr1, in the left part of the cylinder with a cooled working fluid, the gas moves from the connector h to the nozzle about the heat exchanger 12, then through the nozzle and to the nozzle a of the heat exchanger 10. From the output of the heat exchanger 12 (nozzle b), the working fluid enters the freed volume Behind the curtain 9left, but enters through the check valve 14 and channel w_e of the heat exchanger 11 and already at the highest temperature in the cycle.

 By the end of the movement of the curtain-partitions 9left and 9th right, by the end of the process of cooling the working fluid (on the right side of the cylinder) and heating the working fluid (on the left side of the cylinder), the shaft and washers rotate so that the stem begins to slide along the oblique part of the active washer 4right and to do the useful work of rotating the shaft under the action of expanding gas on the left side of the cylinder and a pulling force on the piston into the reduced pressure region on the right side of the cylinder.

 So the working stroke of the piston in the engine. As the stroke progresses, the pressure difference in the left and right parts of the cylinder drops.

At the time of equalization of pressure on the sides of the piston 1, the inclined part of the oblique washer ends, the rod is fixed in the achieved state between the non-inclined parts of the washer, the position of the piston and the volumes in the right and left parts are fixed (although they turn out to be unequal) and the movement of the curtain-partitions begins back side that is accompanied heat transfer from a hot working fluid through a non-return valve 16 to a regenerator 10 and then to a cold working fluid flowing through a non-return valve 16 to a heat exchanger 11 for pre-heating the heated working fluid in a heat exchanger 11, and the process of moving the curtain-partitions is accompanied by the removal of excess heat (not transferred to the regenerator 10) from the cooled working fluid in the heat exchanger pre-cooler 12.

 In the process of heating the working fluid, for example, on the left side of the cylinder, the curtain-partition, moving from right to left (purge Pr1), causes gas to move in the heat exchangers, but the check valve 17 prevents the passage of gas through the on channel of the heat exchanger 12. During the cooling of the working fluid on the left side of the cylinder, the check valve 18 makes the working fluid move through the channel n about the heat exchanger 12 and transfer to this heat exchanger the remnants of waste heat from that heat that was not transferred to the heated body in the regenerator 1 0.

 Check valves 19 and 20 are designed to implement similar procedures when heating and cooling the working fluid on the right side of the cylinder to prevent movement of the working fluid through channel c_p in the heating process of the working fluid on the right side of the cylinder.

 Based on the foregoing, if the thermal resistances of heat exchangers 11 and 12 are correctly calculated, then at the inlet of the adder (mixer), from the fittings and n of the heat exchanger 11, and from the fitting f of the heat exchanger 12, the heat carrier will have the same temperature and this temperature (at the output 22 of the adder ( mixer) 21), will have a temperature (in the case of the cycle shown in figure 2) ΔΤ = 40 degrees lower than the temperature of the coolant at the inlet 23.

 The described device in any technically feasible way, you can constructively pair (connect) with existing electrical generators in order to convert thermal energy into mechanical work and, further, into electrical energy.

 If you make the distance between the two washers slightly larger than is necessary for the cylinder of the selected volume, then on each rod you can place a double-acting piston and place it in the cylinder of the hydraulic drive pump, and the hydraulic drive is often more attractive than the electric drive for reasons of winning the hydraulic motor (hydraulic motor ) for an electric motor according to the specific gravity criterion.

5.3. The implementation of the method according to paragraph 4 of the claims.

 It is known that the approximation of the nature of the thermodynamic processes of compression and expansion to the isothermal process increases the overall efficiency of the heat engine. In the claimed direct and reverse Sterling cycle with controlled heat transfer conditions in the regenerator, the efficiency also increases if the nature of the thermodynamic processes of compression and expansion of the working fluid in the heat engine is close to the isothermal nature of the implementation of thermodynamic processes.

5.4. The implementation of the method according to paragraph 5 of the claims. It is known that the Stirling cycle can be carried out in the opposite direction, i.e. if we rotate the output shaft of a heat engine operating according to the Sterling cycle with an external force in the opposite direction, then the thermodynamic processes in the cycle will occur in the reverse sequence, realizing the heat pump cycle. Taking into account the development trend of heat pumps towards the use of the Sterling heat engine and published information in the source of information: [Tanklevsky V.I., Gruzman PM, Kirillov N.G., Sudar Yu.M. “Decentralized heat supply systems with heat pumps operating on the reverse Sterling cycle”, “Heat-energy-efficient technologies”, Information Bulletin, N 1, St. Petersburg, 1997, pp. 38-40], it is expedient and feasible to obtain heat pumps, cryogenic machines and machines moderate cold with increased (increased) their overall refrigeration coefficient, use the reverse Stirling cycle with controlled heat transfer conditions in the regenerator, where the sequence of operations (alternating stages of the cycle) with a working fluid in thermodynamics cycle is performed in the reverse order.

5.5. The implementation of the method according to paragraph 6 of the claims.

 In a Sterling cycle with controlled heat transfer conditions in the regenerator according to paragraph 1 of the claims, when using a heat engine operating on a direct Sterling cycle, the working fluid spent at the stage of the working stroke in the cycle is sent to the cold regenerator of the heat pump, mated by the temperatures of the cold and hot regenerator, respectively, with the temperatures of the waste heat and the supply heat in the cycle, and the working fluid in the cycle, at the stage of heating the working fluid in the Sterling cycle with controlled heat transfer conditions in the regenerator, it is heated not only by supply heat, but also in a hot radiator of the aforementioned conjugate heat pump.

5.6. Estimation of the specific gravity of the engine performed according to the Stirling cycle with controlled heat transfer conditions in the regenerator.

 The main components of the working units of the engine, made according to the structural diagram shown in FIG. 5 is a piston machine and a regenerator (heat exchanger).

 Solutions of regenerators are known that have a thermal power of more than Rmoshch> 30 000 kW in the volume of one cubic meter with a heat head of not more than dT = 1 degree between heat-exchanging media.

 In the calculations proposed above, a heat engine operating according to the declared cycle will have to pass less than six liters through its working volume per kilogram of air. The diameter of each cylinder (d ^ m) of the engine will amount to no more than 150 mm.

Cylinders around one shaft with oblique washers can accommodate more than one cylinder - for example, eight cylinders. During the operation of such an engine, in one pass of the rod in one direction, the work produced by one rod will amount to L cycle = 0.373 kJ kg, but in one revolution of the shaft, each rod makes a double stroke and returns to its original state.

 Therefore, during the period of rotation of the shaft with the oblique washer, each rod will develop double work (L2), i.e. L2 = 2 * L cycle = 0.373 * 2 = 0.746 kJ / kg.

 And if there are eight units in the engine in the number of cylinders, then for one revolution of the shaft, the total total work (сsumm) will be developed, namely:

 Bsumm = 8 * L2 = 8 * 0.746 = 5.968 kJ.

With a far from maximum engine load, at a frequency (f) of the output shaft speed f = 6000 rpm (100 rpm), all eight cylinders per second will produce a total total work of 596.8 kJ / s, which corresponds to their total power Р Power ⁼ 596.8 kW

 The volume of eight engine cylinders will be equal to Vg = 48 liters. The volume filling factor of the metal is not more than 25%, and the packing density of the engine elements is 0.5.

Hence, the full volume (V _da dz) engine working cycle according to the claimed, will be a value Udaig <100 liters, and its mass will be a value mi <300 kg.

 The volume of heat exchangers (V-GEPLOOBM) must not exceed the value:

VTEIDIOOEM <48 / 30,000 = 0.0016 m ³

 Suppose that not the best heat exchanger is selected, and then, if this heat exchanger is ten times worse than an effective heat exchanger, for example, its volume will be no more than: UTEPLOOBM <16 liters.

The fill factor of the heat exchanger with metal is not more than 25%, this means that its mass will be about t ₂ = 50 kilograms.

Taking into account the volume and mass of the engine, including the heat exchanger, as well as taking into account the necessary refrigerator, the volume of the unit with a power of Rmoshch ⁼ 596.8 kW will be, respectively:

- by volume V = V-GEPLOOBM + Udaig ⁼ 100 + 16 = 1 16 liters,

- by mass M = ni _l + m ₂ = 350 kg.

Then its specific gravity (M _beats ) will be:

M _beats = M / PM power = 350 / 596.8 = 0.586 kg / kW,

 and the specific volume will be:

U _beats = V / Рmoshchn = 1 16 / 596.8 = 0.194 liter / kW.

 Those. we have the following specific engine parameters by weight and volume:

Mood = 0.586 kg / kW, and U _beats = 0.194 liter / kW.

 *** Similar specific parameters of the power plant of the Boeing 737 aircraft (together with the fuel supply) - are the values:

Mood = 2.6 kg / kW, and Uy _D = 1.1 liter / kW.

6. Industrial applicability. The parameters of the claimed cycle shown in FIG. 2 correspond, inter alia, to a device, which may have the name: “Low-grade waste heat converter”.

 Such low-temperature sources of heat include heat transferred to cooling towers of power plants, heat of cooling of powerful compressors and internal combustion engines, parasitic heat of powerful computer centers, heat of metallurgical plants, heat from collectors of exhaust ventilation of the subway, hotels and industrial premises, etc. etc., which is advisable to use, in other words, disposed of with the receipt of mechanical work.

 A distinctive feature of such a converter utilizing low-temperature waste (low-potential) heat is the small temperature difference between the temperature of the heat supply to the converter, made on the basis of a heat engine with an external heat input, and the temperature of the cold tank into which the heat engine is forced to discharge waste heat from the spent working heat body.

 When creating a low-potential heat converter according to the declared thermodynamic cycle, the temperature of the waste heat will be higher than the typical value of the environment (for example, water in a natural reservoir) and this will ensure the operability of such a converter. For reasons of simplifying the design and operation of the low-potential-heat converter to work, conventional compressed air is used as the working fluid in the "Low-potential-heat waste converter", and in order to reduce the specific volumetric characteristics, the average air pressure in the cycle equal to 20.0 MPa was chosen.

 A small change in pressure in the cycle makes it easier to meet the requirements for the reliability of the membrane that separates heat-exchanging media from each other, and a small coefficient of pressure change provides a low temperature difference in the heat supply and removal.

 Actually, the combination of parameters in the cycle may be different, as well as the type of the selected working fluid, for example, air, argon, neon, helium.

 According to the declared thermodynamic cycle, the heat engine is airtight, but there can be no high requirements to the structure for its tightness, in the case of using air as a working fluid, because the working fluid is easily replenished by a small capacity compressor built into the machine.

 Thus, a low-temperature heat converter operating according to the declared thermodynamic cycle allows efficiently processing low-temperature waste heat into mechanical work with an efficiency much higher than the efficiency of the Carnot cycle.

Claims

Claim

1. Forward and reverse Sterling cycle with controlled heat transfer conditions in the regenerator, consisting of thermodynamic processes of heating and compression of the working fluid, expansion and cooling of the working fluid, heat transfer from the spent working fluid to the heated working fluid, characterized in that the cooled working fluid and heated the working fluid is transferred to the regenerator under conditions when both working fluids, heated and cooled, are in a constant volume state.

 2. Forward and reverse Sterling cycle with controlled heat transfer conditions in the regenerator, according to paragraph 1, characterized in that the cooled working fluid and the heated working fluid are transferred to the regenerator of the heat engine under such conditions that when the regenerator is hot-blown, the heat capacity of the material that accumulates heat in the regenerator , set the maximum possible, and with a cold purge of the regenerator, the heat capacity of the heat storage substance in the regenerator sets the minimum possible.

 3. The Sterling cycle with controlled heat transfer conditions in the regenerator of the heat engine, according to paragraph 1, characterized in that

 on each of the four stages of the cycle alternating one after another, two complementary thermodynamic processes are carried out simultaneously with the working fluid in the first and second devices for changing the volume of the working fluid - expanders, namely:

 at the first stage of the cycle, in the first expander, the working fluid is isochorically heated, and in the second expander, the working fluid isochoric is cooled;

 at the second stage of the cycle, in the first expander, the working fluid is expanded to obtain useful work, and in the second expander, the working fluid is compressed and also to obtain useful work; at the third stage of the cycle, the working fluid is isochorically heated in the second expander, and the working fluid isochorically cooled in the first expander;

 at the fourth stage of the cycle, the working fluid is expanded in the second expander to obtain useful work, and the working fluid is compressed in the first expander and also to obtain useful work;

in addition, in isochoric processes of heating / cooling of the working fluid of the first and third stages of the cycle, the working fluids are moved in the regenerator towards each other, preventing their mutual penetration, but create conditions so that from the working fluid previously heated in the previous cycle, into the working the heat that was previously cooled in the previous cycle flowed, provided that the working fluid, previously heated in the previous process of the cycle and passed through the regenerator, was additionally cooled to a lower temperature in the cycle, and the working fluid was previously cooled what was expected in the previous cycle process is additionally heated to the upper temperature in the cycle.

4. Forward and reverse Stirling cycle with controlled heat transfer conditions in the regenerator of the heat engine, according to paragraph 1, paragraph 3, characterized in that at the stages of expansion and compression of the working fluid in the respective expanders expand or compress the working fluid isothermally.

 5. The Sterling cycle with controlled heat transfer conditions in the regenerator of the heat engine, according to paragraph 1, characterized in that in order to obtain a heat pump made according to the reverse Sterling cycle with controlled heat transfer conditions in the regenerator, operations with the working fluid in the heat engine are performed in the reverse order .

 6. The Sterling cycle with controlled heat transfer conditions in the regenerator according to paragraph 1, characterized in that when using a heat engine performing a direct Sterling cycle with controlled heat transfer conditions in the regenerator, the working fluid spent at the stage of the stroke in the cycle is sent to the cold heat pump regenerator, associated with the temperatures of the cold and hot regenerator, respectively, with the temperatures of the waste heat and the supply heat in the cycle, and the working fluid in the cycle being performed, at the stage of heating the worker about the body in the Stirling cycle with controlled heat transfer conditions in the regenerator, they are heated not only by the supply heat, but also in the hot radiator of the aforementioned conjugate heat pump.