RU2191908C2 - System for automatic control of revolution number of stage of gas turbine engine - Google Patents

Abstract

FIELD: automatic control of revolution number of stage of gas turbine engine. SUBSTANCE: system includes electronic control subsystem and electronic hydromechanic actuating subsystem. The last includes driving and feed-dispensing circuits of fuel meter. Feed-dispensing circuit is provided with throttle meter on metering opening of which constant pressure drop is controlled in closed loop by means of slide pressure-drop valve and controlled by said valve piston drive of sloped washer of feeding axial-piston pump. According to analysis upon which all variants of invention are based spontaneous increase of fuel consumption is caused by two main reasons: a) circuit - pressure drop valve _→ piston drive as dynamic link is statically and dynamically unstable because of negative gradient of change of summed forces loading piston of piston pump, namely: force created by pump and force created by return spring, stability of loop with such circuit is provided due to negative hydraulic feedback in it; b) tendency for interruption of hydraulic feedback of loop caused by rigid fixation of slide valve of pressure drop valve unit due to jamming non-filtered foreign particles containing in fuel in variable throttling openings of its very miniaturized command unit. Novelty is dimensional and(or) constructional change of command unit of pressure drop valve and(or) using in piston drive spring unit (instead of return spring) providing more intensive (partial or complete) compensation of negative gradient of change of pump force. EFFECT: improved design allowing to eliminate abnormal modes at operation of system. 8 cl, 10 dwg, 2 tbl

Description

The invention relates to the field of automatic control systems (CAP) of gas turbine engines (GTE). Developed in connection with the specific fact of long-repeating cases of the occurrence of the high-pressure cascade speed (n _in ) of a powerful turbojet engine such as anomalous negative processes such as spontaneous increases in fuel supply, and in cases of the largest increases with an emergency outcome, fluctuations in parameters engine operation in modes close to the idle mode, and a number of other abnormal processes. Such processes are qualified by experts as recurring defects of this CAP. Defects persist due to the great difficulties of knowing a rather complex, nowhere described physical being of one of its circuits, which does not correspond to the constructive solution of individual functional elements of this circuit.

CAP GTDs are known that provide, as described variants of CAP n _vd (hereinafter referred to as variants), automatic regulation by changing the fuel supply and are similar to them in a number of structural and (or) structural features. For example, CAP, for which patent FR 248658 (LUCAS INDUSTRIES LTD) was issued, 01/15/1982, has an identical electronic control part and an executive hydromechanical part, or CAP, for which the copyright certificate SU 566943 is issued (V. Shirshov, V. Yuminov G.), announced on July 23, 1964, is close to them by its dynamic structure and the determining structural features of individual functional elements of the executive part.

 But the closest analogue of the options is CAP, which was originally issued the copyright certificate SU 1090083 (Barsukov A.E. et al.), 01/27/1996, Bull. 3. This analogue, besides having a common structure identical with the options, gives freedom to choose the type of feed pump, which, as will be shown below, is a factor determining the essence of the repeated defects noted above and, accordingly, determines the purpose of the options, which consists in in order to either significantly reduce the number of manifestations of such defects to different degrees, or completely eliminate them.

где индекс "и" - исходный, золотниковый клапан регулирования перепада давлений на дозирующем отверстии дозатора (клапан перепада), поршневой привод наклонной шайбы питающего аксиально-плунжерного насоса (поршневой привод насоса) с эффективной площадью поршня

управляемый клапаном перепада и снабженный пружинным узлом с силой, направленной по оси поршня и передающего штока в сторону наклонной шайбы, насос, связанный каналом питания с топливными баками через двигательный фильтр, при этом задающая часть клапана перепада выполнена для задания перепада давлений

- это вторая формулировка комплекса общих признаков, (18)
которую дополняет первый существенный признак 1-го и 2-го вариантов и следующий второй существенный признак 2-го варианта:
наряду с этим его пружинный узел выполнен из выбранного из ряда 1, ..., n числа параллельно установленных пружин и выбранного из ряда 0, 2, ..., m числа последовательно установленных пружин и при этом снабжен соответственно количеству последних специальными упорами, отвечающими значениям положений поршня поршневого привода насоса (поршня) H₂, ..., H_m, что обеспечивает узлу текущее значение эффективной жесткости с выражением

где H - текущее положение поршня,
H_j+1-(j+1)-e положение поршня,

- жесткость i-й параллельно соединенной пружины,

- жесткость (j+1)-й последовательно соединенной пружины,

- жесткость первой последовательно соединенной пружины,

превышающее значение 4,9 Н/мм, при оптимальном диапазоне значений, определяемом неравенством значений суммы градиентов сил

в выражении которого
N_H(H×n_вдH) - сила от момента на наклонной шайбе насоса, действующая по оси поршневого привода насоса, порежимные значения которой определены экспериментально, где n_вдH - частота вращения каскада высокого давления, приведенная к приводу насоса,
N_П(H) - сила, создаваемая пружинным узлом, вследствие чего ∂N_П(H)/∂H = C_П(H), в связи с чем значения величин s_д,S_по,P_Δ выбраны на основании взаимосвязей, подчиненных следующим условиям

или

где N_H(H) - сила от момента на наклонной шайбе насоса при n_вдН, обеспечивающем максимум суммы сил [N_H(H)+N_П(H)],
при условии, что в диапазоне значений величины s_д выполняется равенство

- это второй комплексный существенный признак 2-го, 4-го, 6-го и 8-го вариантов. (24)
3. Исполнительная часть 3-го варианта включает последовательно. .. комплекс общих признаков, приведенный первой формулировкой (1), командная часть (его) клапана перепада выполнена из двух пар радиально противоположных острокромочных пазов на втулке клапана с суммарной шириной каждой пары пазов, имеющей значение
e=(e^ном±0,1) мм,
где номинал
е^ном≥0,6 мм,
одной острокромочной проточки на золотнике и включена в канал питания параллельно дозирующей части дозатора так, что соединена с каналом подвода топлива в первом подварианте или каналом отвода топлива во втором подварианте, и, соответственно, с отводом топлива по каналам с включенными постоянными дросселями с проливкой Q_с. где индекс с - постоянный, в первом подварианте или с подводом топлива по каналам с включенными постоянными дросселями с проливкой Q_c во втором подварианте, благодаря чему посредством проливочного контроля и установки осевого положения обоих торцов проточки и механической доработки одного из них обеспечивается при работе автоматическая установка жесткой связки двух одинарных гидравлических усилителей с рабочими отверстиями изменяемых дросселей в форме незамкнутых зазоров (с зазорами), текущие значения ширины которых определены значениями ширины h установочных производственных зазоров, состоящими из отмеченного индексом "п" проливочного слагаемого, имеющего, согласно расчетной оценке с возможной погрешностью до ±5%, значение
h_п(Q_и)=[h_п ^нoм(Q_и)±0,02]мм,
где номинал
h_п ^нoм(Q_и)=0,06 мм,
которое получено с помощью проливки по схеме - на входе дроссельный пакет с значением исходной проливки Q_и=200 см³/мин под перепадом давлений 1,0 МПа, проливаемый при входном избыточном давлении (1±0,02) МПа, на выходе - зазор, проливаемый при входном избыточном давлении [(1±0,02)-(0,05±0,01)] МПа, и отмеченного индексом "м" механического слагаемого, имеющего значение, составляющее
у одного из зазоров изменяемых дросселей-
h_м1=(h_м1 ^ном±0,00) мм, (25)
где номинал
h_м1 ^ном={0,00+[h_п ^ном(Q_и)-h_п ^ном(Q)]} мм, (26)
у другого -
h_м2=(h_м2 ^ном±0,02) мм, (27)
где номинал
h_м2 ^ном≥{0,02+[h_п ^ном(Q_и)-h_п ^ном(Q)]} мм, (28)
при этом значение проливки постоянных дросселей выбрано под перепадом давлений 1,0 МПа согласно равенству

где е^ном в мм,
причем фактическое значение проливки Q дроссельного пакета выбрано согласно неравенству
Q≤Q_и,
- это первый комплексный существенный признак 3-го и 4-го вариантов - (29)
его дополняет второй существенный признак 1-го, 3-го, 5-го и 7-го вариантов (17).This goal is achieved in that each version of the system for automatically controlling the speed of a cascade of a gas turbine engine, containing an electronic control part and an electronic hydromechanical executive part, is characterized in that its executive part includes a set of common features and two complex essential features, one or both of which meet the goal, namely:
1. The executive part of the 1st option includes a series of electronic-hydromechanical drive circuit of the fuel metering unit (metering unit) and a hydromechanical power-metering circuit containing sequentially the metering part of the metering unit, the differential pressure control valve on the metering port of the metering unit (differential valve), piston actuator inclined washer of the feed axial plunger pump (piston pump drive), controlled by a differential valve and equipped with a spring assembly with axial force the piston and the transmitting rod towards the inclined washer, a pump connected to the fuel tanks by the power channel through the engine filter, while the master part of the differential valve is made to set the differential pressure,
is the first formulation of a set of common features, (1)
the command part of the differential valve is made of four pairs of radially opposite longitudinal sharp-edged grooves on the valve sleeve with a permissible connection of the intermediate grooves with the total width of each pair of grooves having a value
e = (e ^nom ± 0.1) mm, (2)
where is the face value
^≥ 0.6 mm, (3)
two sharp-edged grooves on the spool and included in the power channel parallel to the metering part of the dispenser, due to which, through flow control and setting the axial position of the four ends of the grooves and mechanical refinement of three of them relative to the lateral edges of the grooves, automatic installation of a rigid bundle of two binary (the name "binary""and" single "are assigned to hydraulic amplifiers in the author's works, for example, in the Technical Information, the purpose of which is described below in the" list n figures and explanations to them. "They relate, respectively, to hydraulic amplifiers with both variable (variable) chokes and one variable, the other constant (constant) choke) hydraulic amplifiers with throttle openings in the form of open gaps (with gaps), current the width values of which are determined by the values of the width h of the installation production gaps, consisting of the pouring term marked by the subscript "p", having, according to the estimated estimate, with a possible error of up to ± 5%,
h _p (Q _and ) = [h ^nom _p (Q _and ) ± 0.02] mm, (4)
where is the face value
^hnom _p (Q _and ) = 0.06 mm, (5)
which is obtained by pouring according to the scheme - at the inlet a throttle package with a nominal value of the initial spill Q _and = 200 cm ³ / min under a pressure drop of 1.0 MPa, spilled at an input overpressure (1 ± 0.02) MPa, at the output - the gap spilled at the input overpressure [(1 ± 0.02) - (0.05 ± 0.01)] MPa, and the mechanical term marked by index m, having a value of
at the input gap of the binary amplifier, which controls the pressure in the chamber of the piston pump drive from the direction of the force of the spring assembly (in the spring chamber)
h _{input m} = (h _{input m} ^nom ± 0.02) mm, (6)
where is the face value
h _{input m} ^nom > {0.05+ [h _n ^nom (Q _and ) -h _n ^nom (Q)]} mm, (7)
at the output gap of the binary amplifier, which controls the pressure in the spring chamber, adopted as the base for pouring settings,
h _out.m = (h _out.m ^nom ± 0.00) mm, (8)
where is the face value
h _vyh.m 0,00+ ^nom = {[h _f ^nom (Q _and) -h _f ^nom (Q)]} mm, (9)
at the input gap of the binary amplifier, which controls the pressure in the chamber of the piston pump drive from the side of the transmitting rod (in the rod chamber)
h ' _inm.m = (h' _inm.m ^nom ± 0.02) mm, (10)
where is the face value
h _'vh.m ^nom = {0,00+ [h _f ^nom (Q _and) -h _f ^nom (Q)]} mm (11)
at the output gap of the binary amplifier controlling the pressure in the rod chamber
h ' _out.m = (h' _out.m ^nom ± 0.02) mm, (12)
where is the face value
h ' _out.m ^nom > {0.05+ [h _p ^nom (Q _and ) -h _p ^nom (Q)]} mm, (13)
moreover, the actual value of the spill Q of the technological throttle package is selected according to the inequality
Q≤Q _and , (14)
- this is the first significant feature of the 1st and 2nd options, (15)
along with this, the spring assembly is made of one spring with a stiffness value
C _P = 4.4 ^+0.5 N / mm (16)
- This is the second essential feature of the 1st, 3rd, 5th and 7th options. (17)
2. The executive part of the 2nd option includes a series of electronic-hydromechanical drive circuit of the fuel metering device (metering device) and a hydromechanical power-metering circuit containing a metering part of the metering unit with a metering hole having an area

where the index “and” is the initial, spool valve for regulating the differential pressure at the metering metering hole (differential valve), the piston drive of the inclined washer of the feed axial plunger pump (piston pump drive) with the effective piston area

controlled by a differential valve and equipped with a spring assembly with a force directed along the axis of the piston and the transfer rod towards the inclined washer, a pump connected by a power channel to the fuel tanks through a motor filter, while the master part of the differential valve is designed to set the differential pressure

is the second formulation of a set of common features, (18)
which is complemented by the first essential feature of the 1st and 2nd options and the following second essential feature of the 2nd option:
along with this, its spring assembly is made of a number of springs installed in parallel from a number 1, ..., n and a number of springs installed in series from a number 0, 2, ..., m, and is equipped with special stops corresponding to the number of the latter, corresponding to the values of the positions of the piston of the piston pump drive (piston) H ₂ , ..., H _m , which provides the node with the current value of effective stiffness with the expression

where H is the current position of the piston,
H _{j + 1} - (j + 1) -e position of the piston,

- stiffness of the i-th parallel connected spring,

- stiffness of the (j + 1) th series-connected spring,

- stiffness of the first spring connected in series,

exceeding a value of 4.9 N / mm, with an optimal range of values determined by the inequality of the sum of the force gradients

in the expression of which
N _H (H × n _vH ) is the force from the moment on the inclined pump washer acting along the axis of the piston pump drive, the mode values of which are determined experimentally, where n _vH is the frequency of rotation of the high pressure cascade reduced to the pump drive,
N _P (H) is the force created by the spring unit, as a result of which ∂N _P (H) / ∂H = C _P (H), and therefore the values of s _d , S _by , P _Δ are selected on the basis of interrelations the following conditions

or

where N _H (H) is the force from the moment on the inclined washer of the pump at n _vdN , providing a maximum of the sum of the forces [N _H (H) + N _P (H)],
provided that in the range of values of s _d the equality

- this is the second complex essential feature of the 2nd, 4th, 6th and 8th options. (24)
3. The executive part of the 3rd option includes sequentially. .. a set of common features, given by the first formulation (1), the command part (of it) of the differential valve is made of two pairs of radially opposite sharp-edged grooves on the valve sleeve with the total width of each pair of grooves having a value
e = (e ^nom ± 0.1) mm
where is the face value
e ^nom ≥0.6 mm
one sharp-edged groove on the spool and is included in the power channel parallel to the metering portion of the dispenser so that it is connected to the fuel supply channel in the first sub-option or the fuel exhaust channel in the second sub-option, and, accordingly, to the fuel drain through channels with constant chokes switched on with Q _s . where index c is constant, in the first sub-variant or with fuel supply through channels with constant chokes turned on with flow Q _c in the second sub-version, due to which automatic installation is provided during operation by pouring control and setting the axial position of both ends of the grooves and mechanical refinement of one rigid connection of two single hydraulic amplifiers with working holes of variable chokes in the form of open gaps (with gaps), the current values of the width of which are determined by by the widths h of the installation production gaps, consisting of the pouring term marked by the subscript “p”, which, according to the calculated estimate, with a possible error of up to ± 5%,
h _p (Q _and ) = [h _p ^no (Q _and ) ± 0.02] mm,
where is the face value
h _f ^nom (Q _u) = 0.06 mm
which is obtained by pouring according to the scheme - at the inlet is a throttle package with the value of the initial spill Q _and = 200 cm ³ / min under a pressure drop of 1.0 MPa, spilled at the input overpressure (1 ± 0.02) MPa, at the output there is a gap spilled at the input overpressure [(1 ± 0.02) - (0.05 ± 0.01)] MPa, and marked by the index “m” of the mechanical term having a value of
one of the gaps of variable chokes -
h _m1 = (h _m1 ^nom ± 0.00) mm, (25)
where is the face value
^SG h = _{m1 0,00+ [h _f ^nom (Q _and) -h _f ^nom (Q)]} mm (26)
another -
h _m2 = (h _m2 ^nom ± 0.02) mm, (27)
where is the face value
h _m2 ^SG ≥ {0,02+ [h _f ^nom (Q _and) -h _f ^nom (Q)]} mm (28)
the value of the spill of constant chokes is selected under a pressure drop of 1.0 MPa according to the equality

where e ^nom in mm
moreover, the actual value of the spill Q of the throttle package is selected according to the inequality
Q≤Q _and ,
- this is the first complex essential feature of the 3rd and 4th options - (29)
it is supplemented by the second essential feature of the 1st, 3rd, 5th and 7th options (17).

 4. The executive part of the 4th option includes sequentially. .. the complex of common features given in the second formulation (18), which complements the first complex essential feature of the 3rd and 4th options (29) and the second complex essential feature of the 2nd, 4th, 6th and 8th options (24).

5. The executive part of the 5th option includes sequentially. .. a set of common features, given by the first formulation (1), the differential part (part) of the differential valve is made of two pairs of radially opposite longitudinal sharp-edged grooves on the valve sleeve with the total width of each pair of grooves having a value
e = (e ^nom ± 0.1) mm
where is the face value
e ^nom ≥0.6 mm
two sharp-edged grooves on the spool and is included in the power channel parallel to the metering part of the metering device, providing, at the same time, either direct connection of the piston drive chamber of the pump from the direction of the spring unit force channel with the outlet from the metering metering hole in the first variant or direct connection of the piston drive chamber pump from the side of the piston drive rod with a channel with an entrance to the dispensing hole of the dispenser in the second variant, due to which, by means of pouring the control and installation of the axial position of two symmetrical ends of the grooves and the mechanical refinement of one of them relative to the lateral edge of the corresponding groove ensures the automatic installation of one binary hydraulic amplifier with throttle openings in the form of open gaps (with gaps) - or gaps adjacent to the ends of the jumper between the grooves in the first sub-variant, or gaps adjacent to the extreme ends of the grooves in the second sub-variant, the current values of the width of which are determined by the values the width h of the installation production gaps, consisting of the pouring term marked by the subscript “p”, having, according to the estimated estimate, with a possible error of up to + 5%,
h _p (Q _and ) = [h _p ^no (Q _and ) ± 0.02] mm,
where is the face value
h _f ^nom (Q _u) = 0.06 mm
which is obtained by pouring according to the scheme - at the inlet is a throttle package with the value of the initial spill Q _and = 200 cm ³ / min under a pressure drop of 1.0 M Pa, spilled at the input overpressure (1 ± 0.02) MPa, at the outlet there is a gap spilled at the input overpressure [(1 ± 0.02) - (0.05 ± 0.01)] MPa, and marked by the index m of the mechanical term having a value of
one of the gaps, for example, the input, as the base for pouring tincture,
h _{input m} = (h _{input m} ^nom ± 0.00) mm, (30)
where is the face value
h _vh.m 0,00+ ^nom = {[h _f ^nom (Q _and) -h _f ^nom (Q)]} mm (31)
another -
h _out.m = (h _out.m ^nom ± 0.02) mm, (32)
where is the face value
h _{output m} ^nom ≥ {0.03+ [h _n ^nom (Q _and ) -h _n ^nom (Q)]} mm, (33)
moreover, the actual value of the spill Q of the technological throttle package is selected according to the inequality
Q≤Q _and ,
- this is the first complex essential feature of the 5th and 6th options, - (34)
it is supplemented by the second essential feature of the 1st, 3rd, 5th and 7th options (17).

 6. The executive part of the 6th option includes sequentially. .. the complex of common features given in the second formulation (18), which complements the first complex essential feature of the 5th and 6th options (34) and the second complex essential feature of the 2nd, 4th, 6th and 8th options (24).

где е^ном в мм,
- это первый комплексный существенный признак 7-го и 8-го вариантов, - (36)
его дополняет второй существенный признак 1-го, 3-го, 5-го и 7-го вариантов (17).7. The executive part of the 7th option includes sequentially. .. a set of common features, given by the first formulation (1), the command part (its) of the differential valve is made of one pair of radially opposite longitudinal sharp-edged grooves on the valve sleeve with a total width of value
e = (e ^nom ± 0.1) mm
where is the face value
e ^nom ≥0.6 mm
one sharp-edged groove on the spool and is included in the power channel parallel to the metering portion of the dispenser so that it is connected to the fuel supply channel in the first and second sub-options or to the fuel drain channel in the third and fourth sub-options and, accordingly, to the fuel exhaust channel with a constant throttle in the first and second sub-options or with a fuel supply channel with a constant throttle on in the third and fourth sub-options to ensure the connection of the piston drive chamber of the pump from the direction of the sludge of the spring assembly with a channel with an outlet from the metering metering hole in the first and third variants or connecting the piston drive chamber of the pump from the transmitting rod side with the entrance to the metering metering hole in the second and fourth variants, which ensures the automatic installation of one single hydraulic amplifier with variable operation a throttle in the form of an open gap, the width of the current values of which are determined spilling Q value _with a fixed orifice pressure drop of 1.0 M and selected according to inequality

where e ^nom in mm
- this is the first complex essential feature of the 7th and 8th options, - (36)
it is supplemented by the second essential feature of the 1st, 3rd, 5th and 7th options (17).

 8. The executive part of the 8th option includes sequentially. .. a set of common features given in the second formulation (18), which complements the first complex essential feature of the 7th and 8th options and the second complex essential feature of the 2nd, 4th, 6th and 8th options (24 )

Figure 1 shows the General block diagram of the regulatory part of the options CAP n _vd (options); figure 2 is a schematic structural diagram of the hydromechanical part of the electronic hydromechanical drive circuit 2a and the hydromechanical feed-metering circuit 2b of the 1st option, 2nd option, excluding its element B, 3rd, 5th and 7th options, excluding their element A, 4th, 6th, 8th options, excluding their elements A and B; figure 3 is a schematic structural diagram of an element A of the 1st and 2nd options; figure 4 - diagram of the forces and gradients of forces created by the pumping unit and the spring unit of the 1st, 3rd, 5th and 7th options (index 1) and sub-options of the 2nd, 4th, 6- of the 8th and 8th variants (serial indices 2 and 3); figure 5 - diagrams of the static mode differential pressure ranges P ₄ - P ₃ on the piston of the piston pump drive applied to the field of the energy pressure differential p ₁ - P ₂ regulated by the differential valve, corresponding to two symmetrical cases (diagram "a") and extremely asymmetric case (diagram "b") of the distribution of the maximum deviations of the dimensions of the installation clearances of hydraulic amplifiers forming the command part of the differential valve, requiring the valve to be replaced with one of the options; in Fig.6 is a schematic diagram of an element A of the circuit of Fig.2, corresponding to the 3rd and 4th variants of the CAP (n _vd ) in Fig. 7 - a schematic circuit diagram of an element A in Figure 2, corresponding to the 5th and 6th embodiments CAP (n _bg) Figure 8 - A schematic circuit diagram of the element 2, corresponding to the 1 st subvariant 7th and Of the 8th version of the CAP (n _vd ) in Fig. 9 is a schematic diagram of the extremely developed element B (spring assembly) of the circuit in Fig. 2; figure 10 is a structural dynamic diagram of the options for CAP (n _VD ) with a detailed presentation of the supply-metering circuit 2b.

Figures 4, 5 and 10 are borrowed from the Technical Information on the results of a physical being study noted above for the repeated defects CAP ( _nd ) of some turbojet engines (hereinafter, for example, “from the Technical Information”). Based on the information contained in this help, options are developed. Research and Technical Reference are executed by the author of the options. The owner of the patent for options possesses the Technical Certificate, which can send its text to interested parties.

 The functional elements contained in figure 2 are shown in some equilibrium working position, which facilitates understanding of their interaction when familiarizing themselves with their work. Functional elements included in the device in Fig.9, shown in the inactive position.

 The options contain (Fig. 1) an electronic digital control part 1, an executive electronic-hydromechanical part 2 and a distribution (distributing fuel along the nozzle contour) part 3. Moreover, the control part 1 includes a program-setting section 1a and a proportional-integral control section 1b, the actuating part 2 sequentially includes an electron-hydromechanical low-inertia (with a time constant of not more than 0.05 s) drive circuit 2a and a hydromechanical inertial (with a time constant of not more than 0.2 s) pita The metering and dosing circuit 2b. In this case, the distribution part 3 can be considered as one of the components of the executive part 2.

The hydromechanical feed-dosing circuit sequentially contains the following functional units:
- a fuel metering device with metering part 4 — as a block of the block diagram in FIG. 1, it is also an adder of differential pressure signals at the metering hole arising from the occurrence of fuel consumption signals and the area of the metering hole;
- spool valve 5 for regulating the differential pressure across the metering hole (differential valve), consisting of a part that calculates the difference signals between the force of the differential pressure at the ends of the spool and the force of the set spring applied to the spool (difference unit), and converts them into spool movement signals ( proportional block), and the command part that converts the movement signals of the spool into pressure signals (the next proportional block);
- piston drive 6 of the inclined washer of the axial plunger pump (piston pump drive), which in combination with the command part of the differential valve 5 forms a high-inertia dynamic link, the essence of which is explained below;
- feeding axial-plunger pump 7 (pump) - it is also an adder of piston drive piston movement signals and own rotor speed signals converting these signals into fuel consumption signals (conditionally equipped with proportional blocks at the inputs of these signals);
- feedback 8 - fuel connection (channel) connecting the fuel outlet from the pump with the entrance to the metering portion of the dispenser.

The structural content of the hydromechanical part of the circuit 2a and the hydromechanical circuit 2b, common to all variants, is as follows (Fig. 2, the positions of the functional devices in Fig. 2 and the corresponding blocks in Fig. 1 are unambiguous):
- the piston drive of the metering part 4 of the dispenser (the output element of the circuit 2a) is the piston executor of 11 commands for connecting two hydraulic amplifiers of the nozzle-damper 12 type, controlled by the common damper 13 of the electrical signal converter of the electronic channel of the circuit 2a and acting on the piston actuator 11 through a spool proportional amplifier converter 14; the corresponding part of the hydromechanical unit is equipped with a spool executor 15 of the function of selecting the operation of the main electronic-hydromechanical and backup hydromechanical CAP (n _vd ) backup control channel of the piston executor 11 - channel 16,
- the dispensing part 4 of the dispenser (in everyday life and hereinafter referred to as the dispensing needle) consists of a hollow, high-sized spool 17 in the sleeve 18, made with non-profile inlets 19 and profile outlets 20 (shown in Fig. 2 in principle), forming with holes 21 on sleeve 18 metering holes; the dispensing needle through the rack and pinion 22 made on it is connected to the analog-digital feedback of the electro-electronic part of the circuit 2a;
- spool valve differential 5 contains a spool 23, a sleeve 24 and a spring 25 with a constant tightening, loading the spool 23 along the axis and sets the pressure drop at its ends; a chamber from the side of the spool end opposite to the end loaded with a spring is connected to the fuel supply line under pressure Р ₁ from the pump to the metering hole, the end-loaded chamber to the spring with the fuel pipe under pressure Р ₁ from the metering hole 2 to the distribution part 3; on the spool 23 and the sleeve 24, structural elements are made that form during operation the command part of the differential valve, consisting of hydraulic amplifiers, the inter-throttle channels of which and the power channels have connections corresponding to the corresponding variant;
- the variable-speed axial-plunger pump assembly 7 contains: a rotor 26 with plungers 27 (9 pieces) equipped with pivotally attached thrust bearings 28 (thrust bearings have spherical bearing surfaces) with a drive spring 29 and a distributor valve 30; an inclined washer 31 having an axis of rotation, a trace of which 32, and provided with a spherical bearing (for spools) surface (the axial section of the washer 31 is not given - the washer has cylindrical protrusions included in the supporting bushings installed and secured in the holes of the housing 33); a separator 34, covering and pressing the thrust bearings 28 to the supporting surface 35 of the inclined washer 31 by the force of a spring 36 acting on the separator through a spherical slider 37; thrust bearings of the rotor 38 (roller) and 39 (needle); a piston drive 40 (aka 6) with a piston 41 (piston sealing devices are shown by shaded rectangles) connected through a transmitting rod 42 with an inclined washer 31 using an earring 43 and axial parts 44 and 45 and equipped with an external or integrated spring assembly 46, force which is directed from the side of the chamber 47, opposite the chamber 48, which includes the rod 42, along the axis of the piston 41 and the rod 42 towards the inclined washer 31; stops of the inclined washer 31 - emphasis (49) of the maximum inclination angle and emphasis 50 of the minimum inclination angle.

) значение модуля градиента силы N_П1, то есть значение жесткости С_П1 пружины.As it is proved in detail in the Technical Information, the physical causes of a dangerous repeating defect CAP n _{in a} certain theater of operations, expressed by a spontaneous increase in fuel supply, are as follows:
1. The command part of the differential valve 5 with a piston actuator 6 controlled by it as a dynamic link is statically and dynamically unstable. The instability of this link is caused by the following reasons:
- with unidirectional action, the force of the spring unit of the piston pump drive (N _P1 ), consisting of one spring, and the force acting on the piston from the pump side (N _H ), have multidirectional gradients (see figa), respectively, with a plus sign and minus
- when the signs of the force gradients N _P1 and N _H differ, the modulus of the gradient of the latter changes by mode and has, in almost the entire range of changes, values that are much higher (see Fig. 4a for the corresponding actual dependences ∂N _P1 / ∂H and

) the value of the modulus of the force gradient N _P1 , that is, the value of the stiffness C _{P1 of the} spring.

где значение слагаемого ∂N_П1/∂H = (4,4-4,9) H/мм.
Это приводит к тому, что при введении активных сигналов воздействия на поршень поршневого привода насоса (изменений перепада давлений на нем) сигналы суммы реактивных сил (N_П1 и N_H) не противодействуют им, как это обычно бывает в устойчивых инерционных гидромеханических звеньях, а помогают. Устойчивость питающе-дозирующего контура 2б в целом обеспечивается действием располагаемой им гидравлической обратной связи, которая дежурно возвращает поршень в каждое исходное положение его неустойчивого равновесия.In the presence of these reasons in the working range, the sum of these gradients, taking into account the production difference of the corresponding aggregates, varies within the limits corresponding to the inequalities

where the value of summand ∂N _P1 / ∂H = (4,4-4,9) H / mm.
This leads to the fact that upon the introduction of active signals of action on the piston of the piston pump drive (changes in pressure drop across it), the signals of the sum of the reactive forces (N _P1 and N _H ) do not counteract them, as is usually the case in stable inertial hydromechanical links, but help . The stability of the supply-metering circuit 2b as a whole is ensured by the action of the hydraulic feedback located by it, which on duty returns the piston to each initial position of its unstable equilibrium.

2. Hydraulic feedback incorporates a differential valve as a link. The command parts of the differential valves are different and are designed so that they ensure reliable operation of the valves. However, it is also possible with the addition of the command part (one of them is described below in the description of the 1st option), in which the spool of the differential valve can be rigidly locked, which will lead to the opening of the hydraulic feedback of link 5-6 and to the piston piston drive the pump from a position of unstable equilibrium when exposed to the smallest axial disturbances. This causes the most dangerous defect CAP n _vd .

Given these reasons, in the wording of the distinguishing parts of the options given in the section of the invention, the essential features cover the following design features:
a) type (binary or single) and the number (two or one) of hydraulic amplifiers of the command part of the differential valve, the main design features and ranges of installation sizes of the throttles of these amplifiers, limited from below (most) or from above, which guarantees the reliability of hydraulic feedback of circuit 2b ,
b) the design of the spring assembly, namely, whether it consists of one spring with stiffness (16), at which the link 5-6 remains unstable, or is solved on the basis of several springs, meeting the general formula (19), which can ensure that the condition (20) link stability 5-6,
c) a combination of signs “a” and “b”, which determines the degree to which the goal of options is provided.

In more detail from certain points of view than in the section "essence of the invention", and clearly significant features of the options CAP n _{vd are} described below:
The 1st option contains a differential valve 5 (Fig. 2, element A and Fig. 3), the command part of which is designed to ensure, during operation, the automatic installation of two binary hydraulic amplifiers U ₄₇ (1) and U ₄₈ (1) of the type spool ", separately controlling the pressure, respectively, in the chambers 47 and 48 of the piston pump drive. From the elements B1, B2 and B3 and the views along the arrows G1, G2 and G3 shown in FIG. 3, it is seen how the throttle openings of these amplifiers are formed and what shape they have. The specified minimum value of the nominal length of the holes (see (3)) e ^nom / 2 = 0.3 mm, the specified minimum value of the nominal width (see (4), ..., (13)) -
h _p ^nom + (h _{input m} ^nom + h _{output m} ^nom ) / 2 = 0.085 mm [or h _p ^nom + (h _{input m} ^'nom + h _{output m} ^{' nom} ) / 2 = 0.085 mm]. Given that the value of their length is more than 3 times the value of the width, the latter being about 0.1 mm, throttle openings can be considered as open gaps (gaps). The same pattern continues and at higher values of these dimensions in predetermined ranges, since according to preliminary calculations their upper limits are most favorable f ^nom / 2 = (0.6-0.8) mm and h _f ^nom + (h _{Rin. m} ^nom + h _{output m} ^nom ) / 2 = (0.14 - 0.16) mm. A further increase in the values of these sizes will be undesirable and then unacceptable due to a corresponding increase in the power consumption of the differential part of the differential valve, which is connected in parallel with the fuel consumption through the metering hole of the metering needle, which will lead to an undesirable or unacceptable error in the main fuel metering. But preserving the above minimum values of these dimensions or close to them for the size f ^nom / 2 is undesirable, as for the size of h _f ^nom + (h + h _vh.m ^prefecture _vyh.m ^nom) / 2 is not allowed. This is explained by the information contained in the following excerpt from the conclusions of the Technical Information, which is devoted to the results of a case study that is structurally identical to the 1st option, but having practically the smallest gap sizes in the given ranges:
<< b) the order of the working values of the areas of the throttle openings (gaps) of the amplifiers is determined by the values of the two installation components of the width of the gaps - provided by machining ... and provided by pouring ...;
...

d) pouring components of the installation width of the gaps determine the zero coordinates of the mechanical components, while the components of both types have the values given in table 1,
the distribution of the total values of the installation gap widths between the input and output gaps of each of the amplifiers shows evenly that the values of the distributed gaps of the amplifier controlling the pressure in the spring chamber are 0.085 ± 0.03 mm (instability ± 35%), the gaps of the amplifier controlling the pressure in the rod chamber, 0.085 ± 0.04 mm (instability ± 47%);
e) depending on the signs and values of the maximum deviations of the installation width of the gaps, symmetrical and to some extent asymmetric cases of tandem [bundles] of amplifiers are distinguished (given in square brackets as part of this excerpt refers to the description of the invention): in symmetrical cases, all sizes of the gaps have the same sign and value limit deviations;
e) in connection with the erosion method of manufacturing grooves on the valve sleeve, the fixed long sides of the gaps have circular transitions in width with a radius of 0.1 mm [Fig. 3, element D]; this, with widths of less than (0.07-0.08) mm, creates a pronounced wedge-shaped contour of the gaps of the movable long sides;
g) the edges along the entire perimeter of the gaps are technologically performed only practically sharp - in fact they have a blunting, the values of which are 0.02-0.03 mm; the blunting profile in the section is close to a chamfer with some ovality, the blunting surface is rather rough;
h) the valve [differential valve] uses a working fluid (fuel), filtered only by a motor filter with a filtration purity of 16 microns.

2.2. Based on the information set out in clause 2.1, a version is developed as follows:
With certain combinations of values and signs of maximum deviations of the tandem gap widths of the amplifiers, the latter become sharply differently effective. As a result, during the automatic installation of the throttling conditions necessary to ensure a balance of forces on the piston of the pump unit drive. the width of one of the gaps (critical) becomes equal to or less than 16 μm (I). Such a gap acquires a stable susceptibility to clogging, up to a jam in it of unfiltered foreign particles (II) stably contained in the used fuel. Jamming is facilitated by wedge-shaped sections of the contour of the gap (expanding the exposure range) and blunting of the edges (particles can be tightly sandwiched between them). Particle jam completely interrupts valve operation. In this case, the feeding-dosing circuit and CAP n _vd are generally opened and, due to the internal static instability of the circuit, when disturbances corresponding in sign occur, an emission [increase in fuel supply to the engine] with a positive or negative sign (decrease in fuel consumption) occurs. Consequently, emissions are a natural phenomenon determined by the probability of the following independent events occurring: events of industrial origin — I, events of operational origin — level of foreign particles in a unit volume of fuel used — II, and events of working origin — coincidence of the closure of the critical gap with the moment of passage through it a foreign particle.

A calculation and graphical method has been introduced, which for static modes, the version is confirmed within the following boundaries (the results are presented in diagrams):
- an amplifier with a higher value of the sum of the installation areas of gaps has higher efficiency;
- emissions with a positive sign may occur as a result of jamming of foreign particles in the output gap of the amplifier, which controls the pressure in the rod chamber [in the gap h ' _o ];
- the lowest probability of occurrence of emissions when the values of the installation clearances are symmetrical; among these cases, the cases with maximum negative values of the maximum deviations are most susceptible to clogging with foreign particles [in figure 5, the diagrams “a” correspond to these cases - the last case on the left],
- the highest probability of occurrence of emissions in the cases when the following maximum installation values of limit deviations occur: at the gaps of the amplifier controlling the pressure in the spring chamber ...- at the input to plus, at the output to minus, at the gaps of the amplifier controlling pressure in the rod the camera. .. - at the input to plus, at the output to minus [in FIG. 5, the diagram “b” corresponds to this case], while in addition, the width of the grooves on the valve sleeve belonging to the amplifier ... is maximum (0.35 mm): >>
The pressure differences P ₄ - P ₃ indicated in FIG. 5 correspond to the following modes: MG - low gas, CR ₁₁ - cruising at an altitude of 11 km, KP ₅ - at an altitude of 5 km, MP - maximum, PD - emergency, PP - reverse, 2 - the conditional reference mode, indicated by a dotted line - the static border of the range of 100% possible clogging of the gap h ' _out (range - towards the increase of the gap h _out ), indicated by the dash-dot line - the static border of the partial approximately 20% possible clogging of the gap h' _out .

The possibility of the formation of a critical working gap h ' _out <16 μm in the Technical Reference is shown by other diagrams. In addition, it shows that the formation of a critical working gap h _o contribute to dynamic disturbances. It also provides an interpretation of the repeating defect of the real CAP n _{in the} theater of operation "fluctuations in engine operation parameters in regimes close to the low-gas mode" noted in the introductory section.

The cited excerpt from the Technical Information, outlining the mechanism of occurrence of the second main cause of the defect “spontaneous increase in fuel supply to the engine” confirms that the wording of the first essential feature of the 1st and 2nd options (15) is correct: it provides such an increase in the values of the optimal width of the installation production gaps , which eliminates the possibility of the formation of a critical gap h ' _out .

 But this feature in the 1st embodiment is combined with the solution of the second essential feature (16), which does not provide condition (20) for link stability 5-6.

The 2nd option contains a differential valve 5, the command part of which is made exactly the same as that of the differential valve of the 1st variant. But unlike the 1st option, the 2nd option contains its solution of the second essential feature - the spring assembly of the piston drive of the pump 6 (element B in figure 2), which meets the fulfillment of condition (20) of stability of the link 5-6. This solution is described in detail below. It will be seen from his description that a solution can have a number of constructive sub-options. All of them are characterized by an increase in all operating modes of the required values of the differential pressure of the command P ₄ -P ₃ (figure 3). In this case, only the most constructively developed sub-option can be used while maintaining the inherent in the first solution of the second essential feature value of the available energy pressure difference P ₁ -P ₂ , which is controlled by a differential valve 5 and the increase of which is technologically undesirable. Therefore, all variants with a spring assembly corresponding to formula (19) are focused on the use, first of all, of just one of its developed sub-options, requiring minimal correction (possibly without it) or effective piston area _Spo (see (21)) a piston pump drive to lower the pressure drop P ₄ -P ₃ or increase the energy drop P ₁ -P ₂ = P _Δ (see (22)) and lower it to maintain the fuel consumption law of the current area of the metering hole s _d (see (2 )).

The 3rd and 4th options contain a differential valve 5 (Fig.6), the command part of which consists, as in the 1st and 2nd options, of two hydraulic amplifiers that separately control the chambers 47 and 48 of the piston drive of the pump 40 ( Fig. 2), but unlike the 1st and 2nd options, the amplifiers are not binary, but single - at ₄₇ (3) and ₄₈ (3). The options are presented in two sub-options: for the 1st sub-option (Fig. 6a), constant chokes (preferably chokes) are output, and in the 2nd (Fig. 6b) are input. The 1st sub-option seems to be more preferable, since in it the jets flowing from the gaps are less susceptible to the influence of rotation: they flow from the valve. This increases the operational stability of the valve.

The advantage of the 3rd and 4th options over the 1st and 2nd, determined by the principle of the differential valve command part (single amplifiers), is the weakening of the susceptibility to the formation of a subordinate critical gap in one of the amplifiers - a feature that is especially unfavorable when choosing small values denominations of installation clearances. The interdependence of the gaps of different amplifiers in these variants is preserved, but it is much weaker than that of the first and second variants, since it is determined by a fairly stable installation spacing Q _{c of} constant chokes.

 The difference between the 4th option and the 3rd one is the same as that of the 2nd option from the 1st one - it contains the same spring unit of the pump 6 piston drive as the 2nd option, which ensures the fulfillment of stability conditions (20) link 5-6.

 The 5th and 6th options contain a differential valve 5 (Fig.7), the command part of which is solved on the basis of one binary amplifier. Both options, as well as the 3rd and 4th, are represented by sub-options. In the 1st subvariant (Fig. 7a), the amplifier controls the rod chamber of the piston drive, in the 2nd (Fig. 7b) - spring. Accordingly, the uncontrolled chambers opposite these chambers are directly connected - in the 1st sub-variant with the output 20 (Fig. 2) from the metering hole of the metering needle 17, in the 2nd - with the input 19.

 In the design of the command part of the differential valve of the 5th and 6th options, as can be seen from Fig.7, the disadvantages of the differential valve of the 1st and 2nd options are taken into account. They exclude the possibility of the predominance of the effectiveness of one of the rigidly coupled amplifiers over the other with the formation of a critical gap in the latter.

 The difference between the 6th option from the 5th one is the same as that of the 2nd from the 1st one - it contains the same spring unit of the piston drive of the pump 6 as the 2nd version, which ensures the fulfillment of condition (20) of link stability 5-6.

 In the 7th and 8th variants, the command part of the differential valve (Fig. 8) is solved on the basis of one single amplifier. The scheme of the command part, as in a number of previous options, can be represented by sub-options - there are 4. They are determined by the choice of the controlled chamber of the piston pump drive and the installation location of the variable and constant chokes in the amplifier, which has little effect on the working and quality features of the sub-options. On Fig presents a sub-option, which is controlled by the stock camera, and a constant throttle is the output.

 The advantage of the 7th and 8th options is the structural design and tuning simplicity (no need for flowing and mechanical adjustment) of the command part.

 The difference of the 8th option from the 7th one is the same as that of the 2nd option from the 1st one - it contains the same spring unit of the pump 6 piston drive as the 2nd option, which ensures the fulfillment of stability condition (20) link 5-6.

One of the components of the first essential feature of each of the above described options is the indication of the lower limits of the ranges of assignment of the values of dimensional quantities e ^nom , h _m ^nom at the 1st, 2nd, 3rd, 4th, 5th and 6th variants and Q _c in the 7th and 8th variants [see: (3) in composition (2), (7) in composition (6) when accounting for (9) in composition (8), (13) in composition ( 12) when accounting for (11) in composition (10) for the 1st and 2nd options, (28) in composition (27) when accounting for (26) in composition (25) for the 3rd and 4th options, (33) in composition (32) when accounting for (31) in composition (30) for the 5th and 6th options and (35) for the 7th and 8th options] and knitting _and the upper limit Q (14) set the value range Q, which determines the value of h _n (Q) (see. (4)). The purpose of all these instructions is to prevent the possibility of formation during operation of the corresponding differential valves of critical values of the width of the throttle clearances. When calculating the above lower limits, the values of the working offsets of the moving sides of the gaps from the positions of the symmetric distribution between the rigidly coupled hydraulic amplifiers and inside them the minimum total installation widths of the gaps, as well as the possible error of such calculations, which is about 0.010-0.015 mm, are taken into account.

 As for the indication of the appropriate direction of the choice of the values of Q, (see (14)), it is due to the desire to reduce the instability of the total width of the installation gaps by reducing the instability of their pouring component, in which the value of Q is proportional to both the nominal and the maximum deviation. Moreover, inequality (14) also allows the complete exclusion of the pouring component (Q = 0), which is usually not used in practice in cases of values of the installation gap widths of a few tenths of a millimeter.

 The presence of the above-described component of the first essential feature of the 1st, 2nd, 3rd, 4th, 5th, 6th, 7th and 8th options provides these options with equal high reliability due to this feature.

 With equal values of the installation clearances, the options differ in the values of their working changes. They are minimal in cases of the command part with two binary amplifiers, maximum in cases of the command part with one single amplifier. But, as shown in the Technical Information, analytically this difference practically does not affect any quality indicators. Therefore, in this respect they are also equivalent.

 Since the options are equivalent for the two working characteristics noted above, it is preferable to use, as simpler, differential valves with a command part containing one each - a hydraulic amplifier, and not two. In this case, binary amplifiers appear to be more preferable than single ones, since they exclude the use of throttle packages, which introduces a certain complication in the corresponding unit. As a result, the most preferable for application are variants of the 5th and 6th, containing the command part of the differential valve, based on one binary amplifier.

 A calculated comparison of the options according to the power consumption level of the command part of the differential valve, which is reflected in the error of the main fuel dosing, was not performed. To finally decide which option is better on this basis, it is necessary to make such a comparison when choosing an option for implementation.

 It should also be noted that the reliability of the normal operation of the differential valve is reduced by its spool-piston solution. With it, foreign particles less than 16 microns that are not filtered by the engine filter and contained in the fuel can fall into the circular wedge-shaped grooves in the cross section created by the technologically inevitable blunting of the formally sharp end and grooved edges of the spool (Fig. 7, element B). In this regard, during working movements of the spool, a certain effect of its jamming is likely. The probability of complete jamming is extremely small, but mashing and delays can occur in the counted number of times. Foreign particles are most likely to enter the wall grooves of those bridges between the channels where there is a pressure drop and the fuel movement directed into the grooves, which seeps through the gap between the surfaces of the spool and the sleeve. In all of the diagrams shown, variants of spool valve pairs of the differential valve (Fig. 3, Fig. 6, Fig. 7, Fig. 8) such jumpers are marked with small arrows indicating the direction of leakage. All of these schemes, within the framework of the wording of the options, are solved in such a way that they have only one loaded jumper. But within this framework, circuit solutions are possible having more than one loaded jumper. However, such a difference should be taken into account by another patentable solution. Nevertheless, it is shown that according to the reliability criterion described in this paragraph, all options can be implemented equivalent.

 The solution of the 2nd essential feature of the options that meets the objective (20) - the second solution of the spring assembly, can have a number of design sub-options. The choice of the optimal sub-option for the intended specific case should be carried out by a thorough design study.

It is assumed that the maximum value of the gradient ∂N _П / ∂H (it is determined from inequalities (37)) specified in the Technical Reference describing the specific practical case is determined quite accurately and the margin of its excess modulo by 3 to 5 -and N / mm to ensure the guaranteed sum in operation ∂N _H / ∂H + ∂N _П / ∂H> 0 will be sufficient.

There is a sub-option of the second solution of the spring assembly, structurally simple enough - one spring with a stiffness of 37 N / mm. To the case of such a node in FIG. 4, dependence (19) with row i = 1 and row j = 0 corresponds. The disadvantages of this option:
(1) In order to fit it into the diameter of the piston drive chambers, which is widely used in practice (D _by = 40 mm - see Fig. 3), a spring is required for which, with an external diameter of 38 mm, the maximum working height is 65 mm. This is a very bulky spring - a sub-option will require increased dimensions of the corresponding units in length.

(2) As can be seen from the graph of the force N _H + N _P2 (Fig. 4), the subvariant has a very high maximum value of 1000 N / mm and a corresponding maximum value of the pressure drop across the piston - P ₄ -P ₃ = 1.08 MPa with the available preferred energy pressure difference P ₁ -P ₂ = 0.784 MPa. It will also require some appropriate rework.

 Another sub-variant of the spring assembly, one of the most complex (Fig. 9 is a section along the axis of the rod 42 in Fig. 2).

С учетом известных формул для определения суммарной жесткости блока последовательных и параллельных пружин текущая суммарная жесткость этого узла имеет следующее выражение

включена при Н>H_j+1), которому на фиг.4б соответствует графическая зависимость ∂N_П2(H)/∂H ступенчатого характера. Эта зависимость в большей части диапазона координаты Н компенсирует отрицательный градиент ∂N_H(H)/∂H силы насоса с небольшими избытками, что показывает ее симметричное относительно оси Н пунктирное изображение. Зависимости ∂N_П3(H)/∂H отвечают параметры концентричных пружин, приведенные в таблице 2.The assembly contains: two symmetrically located identical blocks of series-connected concentric compression springs 55, 56, 57 and 58 with an axis parallel to the axis of the rod and in the same plane as it, contact-contact bushings 59, 60 and 61, which provide communication between the springs and the blocks provided with annular thrust projections E ₁ , E ₂ and E ₃ with which all the bushings with extremely compressed springs abut against the flat end face T of the corresponding support plate 62; the plates 62 are interconnected by a rod 63 provided in the middle of a spherical concave supporting surface, which it rests on the sphere of the convex centering sleeve 64, and through it to the ends of the carrier pin 65 fixed in the rod 42. In addition, a parallel spring 66 is included in the assembly installed in the spring chamber 47 of the piston pump drive. The springs 55, 56, 57, 58 and 66 have, respectively, the following stiffness values:

Given the known formulas for determining the total stiffness of a block of sequential and parallel springs, the current total stiffness of this node has the following expression

included for H> H _{j + 1} ), which in Fig. 4b corresponds to a graphical dependence ∂N _P2 (H) / ∂H of a stepwise nature. This dependence in most of the coordinate range H compensates for the negative gradient ∂N _H (H) / ∂H of the pump force with slight excesses, which shows its dotted image symmetrical about the axis H. Depending ∂N _P3 (H) / ∂H parameters correspond concentric springs listed in Table 2.

In the table. 2 is indicated: D _int - inner diameter, d - wire diameter, n - number of working turns, H _Pru - spring height. The material of the springs is 51XFA.

Advantages of the second sub-option:
(1) The spring 66 fits into the dimensions of the piston actuator, for which the calculation was oriented, with insignificant changes in the design of its parts, unlike the spring of the first sub-variant. True, the ultimate force created by this spring is also significant - 44 N. When loading the piston 41 directly, this is undesirable. A close sub-variant of the spring assembly is possible, in which a spring of the same purpose loads the inclined washer in the lower part relative to the axis of rotation, creating a moment equal to the moment of the spring 66. The maximum pressure drop is P ₄ -P ₃ = 0.743 MPa, that is, it is small stock, but in the framework of the energy differential pressure P ₁ -P ₂ fits.

(2) The arrangement of parallel spring blocks is most optimal: the point of application of the total force of the blocks can be provided coinciding with the point of application of the force N _Н - this is favorable for the piston drive to operate, since it will be unloaded as much as possible from an unforeseen moment, which will cause some inevitable small divergence of directions forces N _H and N _P.

 (3) The sub-option will fit into the existing limit dimensions of the unit. Certain alterations of a number of parts, body casting will be required, but in the aggregate regarding the total aggregate complexity, it is insignificant - no more than 5%.

 The disadvantage of the 2nd sub-option in comparison with the 1st is that it is much more complicated than the latter.

Сумматор имеет 3 входа и выход, что отражено стрелками: вход сигналов площади дозирующего отверстия с порядковым номером преобразования 4.1, вход сигналов обратной связи с порядковым номером конечного преобразования 4.2, реализуемых цепью блоков 4-->5-->6-->7-->8-->4 и приходящих со знаком минус, что обеспечивает сумматору функцию, идентичную функции блока разности, располагающего инвертирующим входом, необходимым для замыкания контура. 3-й вход - это вход сигналов

вводимых первично в звено 7 (в насос) в виде сигналов частоты вращения. Канал введения сигналов на 3-й вход приведен пунктиром. Анализ показал очень малое влияние 3-го входа на динамику CAP n_вд выражаемое изменением запаса устойчивости по фазе на более чем на 0,5 градуса, поэтому 3-й канал ввода сигналов в сумматор 4 можно в учет не принимать.The dynamic structure of circuit 2b of all variants is shown in FIG. 10. The sequence numbers of the amplification factors of its actual and conditional links (the latter are conversion links) coincide with the sequence numbers of the corresponding blocks in the block diagram (Fig. 1). At the input, the dynamic structure has a hydromechanical adder (dosing part of the dosing needle) that performs the operation of summing the differential pressure signals at the dosing hole

The adder has 3 inputs and an output, which is reflected by arrows: the input of the metering hole area signals with the serial number of the transformation 4.1, the input of the feedback signals with the serial number of the final transformation 4.2, implemented by the chain of blocks 4 -> 5 -> 6 -> 7- -> 8 -> 4 and those coming with a minus sign, which provides the adder with a function identical to the function of the difference block, which has an inverting input necessary to close the circuit. 3rd input is a signal input

introduced primarily in the link 7 (in the pump) in the form of speed signals. The channel for introducing signals to the 3rd input is shown by a dotted line. Analysis showed very little impact of the 3rd input dynamics CAP n _bg expresses the change in the stability margin in phase by more than 0.5 degrees, so the third input channel signals in the adder 4 can not take into account.

В ней коэффициент усиления -

постоянная времени -

где

Существо входящих в выражения (39), . . . , (43) величин следующее:

см. на фиг.3,

коэффициент объемной гидравлической проводимости, где μ - коэффициент расхода, ρ - плотность рабочего тела (топлива),
C_Σ = C_П(H)+C_H(H).
Как уже известно, в случаях вариантов 1, 3, 5, 7 имеет место сумма (44), причем
|C_H(H)|>|C_П(H)|.
При этом цепь звеньев 6-->7 статически неустойчива и передаточная функция (39) приобретает вид

то есть вид передаточной функции динамически неустойчивого звена 1-го порядка.Circuit 2b is based on an inertial link 6 having a transfer function

The gain in it is

time constant -

Where

Being in the expressions (39),. . . , (43) the quantities are as follows:

see figure 3,

volumetric hydraulic conductivity coefficient, where μ is the flow coefficient, ρ is the density of the working fluid (fuel),
C _Σ = C _P (H) + C _H (H).
As is already known, in the cases of options 1, 3, 5, 7, the sum (44) holds, and
| C _H (H) |> | C _P (H) |.
Moreover, the chain of links 6 -> 7 is statically unstable and the transfer function (39) takes the form

that is, the form of the transfer function of a dynamically unstable 1st order link.

In the cases of options 2, 4, 6, 8, and 9, in which inequality is ensured (see (20))
| C _H (H) | <| C _P (H) |.
link 6 has a transfer function of the form (39).

где К_ос=К_4.2 К₅ К₇ K₈.Accordingly, the general transfer function of circuit 2b for all variants is the transfer function

where K _OS = K _4.2 K ₅ K ₇ K ₈ .

а контур 2б имеет передаточную функцию

В первом приближении все изложенное останется справедливо для всех вариантов. Действительно, в них целесообразное повышение произведения eh'_вх(см (41) и (43)) не превысит 4-х-5-и раз - значения постоянной времени Т₆ по-прежнему останутся очень большими (варианты 1, 3, 5, 7), а в сочетании с существенным признаком (19), при реализации которого вынужденно будут согласно (20) стремиться к минимуму его повышения, это тем более останется справедливо (варианты 2, 4, 6, 8 и 9).The Technical Reference also shows that the values of the time constant T ₆ (41) are ten times greater than unity, which can be neglected regardless of its sign, and therefore link 6 is a practical integrator with a transfer function

and circuit 2b has a transfer function

In a first approximation, all of the foregoing will remain true for all options. Indeed, in them, a reasonable increase in the product eh ' _in (see (41) and (43)) will not exceed 4-5-5 times - the values of the time constant T ₆ will still remain very large (options 1, 3, 5, 7), and in combination with the essential feature (19), during the implementation of which, according to (20), they will be forced to strive to minimize its increase, this will be all the more true (options 2, 4, 6, 8, and 9).

Values of the time constant T _2b depending on the mode are in the range from 0.010 to 0.150 s, its minimum values occur in the regimes close to the low gas mode, and the maximum values occur in the regimes with high values of n _id and fuel consumption.

When implementing the options, as one of the restrictive conditions for choosing the product value eh ' _in, one should accept the condition of insignificant exceeding of such a level of T _2b values, for example, to T _2b = 0.2 s, no more. Moreover, in modes with medium and high fuel consumption, all options in the absence of abnormal situations have good dynamic quality, for example, in cruising mode at
11 km altitude phase stability margin is from 46 to 48 degrees, the amplitude - of about 800%, 1st cutoff frequency - about 2.5 rad / s at a dominant in many CAP n values _vd 1st cutoff frequency of 1, 5 to 2.0 rad / s.

The options work is the process of automatically recovering a given value of n _vd when deviating from it under the influence of defining external and internal disturbances. Since the options for the occurrence of spontaneous increases in fuel consumption and other abnormal negative processes are largely weakened or completely eliminated, we consider the case of automatic restoration of a given value of n _vd when it deviates, plus Δn _vd from the influence of an external disturbance.

проходит, претерпевая предусмотренные операционные преобразования, через блоки секции 1б, поступая на ее выход уже в форме ПИ - сигнала, и затем, как отклонение в угловом выражении поворота валика реечной передачи 22 (фиг.2), через блоки контура 2а, приобретая в нем некоторую инерционность.The deviation plus Δn _vd , acquiring a code expression in the proportional converter unit with the transmission coefficient K _{p → ξ} (Fig. 10), is fed to the inverting input of the central difference unit and passed to the output with the changed sign (minus sign). Further, with the same sign, it is as a deviation

passes, undergoing the provided operational transformations, through the blocks of section 1b, arriving at its output already in the form of a PI signal, and then, as a deviation in the angular expression of the rotation of the rack and pinion roller 22 (Fig. 2), through the blocks of the circuit 2a, acquiring in it some inertia.

The output signal of the circuit 2a is the movement ΔL _{AND of the} spool 17 of the metering needle 4 to the side, providing a decrease in the area ΔS _{AND of the} metering hole, respectively, by ΔS _AND (Fig.2 to the left). Here, the present argument requires the use of the following conditional assumption, which will lead to the correct final result: moving the dosing needle, which changes the area of the dosing hole by ΔS _AND , does not lead to a change in fuel consumption through the needle, since the latter can only happen due to a change in the capacity of the feed pump, which has not yet occurred, therefore, it leads to a transformation of the displacement ΔL _AND entirely to a change in the pressure drop across the metering hole plus ΔP _Δ1 , which occurs when Variable fuel consumption. The unbalanced signal plus ΔP _Δ1 loads the differential valve spool 23 and the spring 25 through it. The spring 25 compresses accordingly - the differential valve spool with a transmission coefficient K ₅ moves Δh to the side (in the direction of the arrow in Fig. 3), which ensures hydraulic throttling by changing amplifiers (h _in and h ' _out in negative, h' _in and h _out in plus) the influx of the working fluid into the rod cavity 48 and the outflow from the spring cavity 47 of the piston actuator 40 (6), which is a link close to the hydraulic amplifiers integral. In this case, the piston 41 with the rod 42 is moved by ΔН to the side, which provides a decrease in the installation angle α _{w of the} inclined washer 31, which causes a decrease in fuel consumption by ΔG _H , which is transmitted to the metering hole 20 of the metering needle 4 as ΔG _AND .

The result of this reduction is the conversion of the signal at the metering hole into a differential pressure signal minus ΔP _Δ3 . The signals plus ΔP _Δ1 and minus ΔP _Δ3 are _summarized in the initial output signal ΔP _{Δ2 of the} dispensing part of dispenser 4. But since the described process is continuous and actually does not occur sequentially, as described, but at the same time, the signal ΔP _Δ3 , as follows from set forth and can be seen from figure 10, for the circuit 2b as a whole is the output that completely determines the output signal of the fuel consumption ΔG _{AND the} regulatory part CAP n _vd which passes through the distribution part with the transmission coefficient K ₃ to the engine.

The described process, which causes a restoration of the initial value of n _vd , a decrease in fuel consumption by ΔG _I , provides either the complete elimination of the deviation plus Δn _vd (aperiodic process), or the formation of a reduced deviation of the opposite sign (oscillatory process) with the subsequent repetition (one to two times) of the described process with the opposite sign by signals until the deviation is almost completely eliminated.

 For options in which complication of the design of the spring assembly is provided, we briefly describe the operation of its more complex sub-option (Fig. 9).

 When the position of the inclined washer 31 (figure 2) on the emphasis 50 of the minimum performance of the spring 55, ..., 58 are maximally compressed, contact-contact sleeve 59,. .., 61 are supported on the end face T of the plate 62. After removing the inclined washer from the stop 50, as the H coordinate increases from point 1 (Fig. 4), the spring force 55 stretches and decreases. At point 2, its force becomes equal to the maximum spring force 56. With a further increase in the coordinate H, the spring force 55 continues to decrease. Maintaining a balance of forces between successive springs, weakening begins by stretching the compression spring 56, which removes the contact-contact sleeve 59 from the stop - the successive springs 55 and 56 work together. Similarly, at point 3, the engagement of the spring 57 and the removal of the support-contact sleeve 60 from the stop starts, at point 4 — the activation of the spring 58 and the removal of the support of the contact-contact sleeve 61 from the stop until the calculated spring extension at point 5, which is shown in FIG. 9.

In conclusion, the importance of the design information introduced by the options for the development of the CAP gas turbine engine, in which it will be necessary to use high-flow axial-plunger pumps, should be emphasized. In addition to these pump specialists, it is advisable to conduct exploratory work at a fundamental level to reduce the level of the ∂N _H / ∂H dependency in each specific case.

Claims (5)

1. A system for automatically controlling the speed of a cascade of a gas turbine engine, comprising an electronic control part and an electronic-hydromechanical actuator, characterized in that the actuator includes a series of electronic-hydromechanical drive circuit fuel dispenser (dispenser) and a hydromechanical power-metering circuit containing sequentially metering part of the metering valve, differential pressure control valve on the metering metering hole (valve turnip), piston drive inclined washers of the feed axial plunger pump (piston pump drive), controlled by the differential valve and equipped with a spring assembly with a force directed along the axis of the piston and the transmission rod towards the inclined washer, a pump connected to the fuel supply channel through the engine a filter, while the differential valve master portion is configured to set the differential pressure, the differential valve command part is made of four pairs of radially opposite longitudinal sharp-edged grooves on the sleeve valve performance with an acceptable intermediate compound with a total width of the grooves of each pair of grooves having a value of e = (e ^nom ± 0,1) mm, where e denomination ^prefecture ≥0,6 mm, two sharp-edged grooves on the spool and is included in the supply channel parallel to the dispensing portion dispenser, due to which, through pouring control and setting the axial position of the four ends of the grooves and mechanical refinement of three of them relative to the lateral edges of the grooves, automatic installation of a rigid bundle of two binary hydraulic amplifiers is ensured during operation th with throttle openings in the form of open gaps (with gaps), the current values of the width of which are determined by the values of the width h of the installation production gaps, consisting of the pouring term marked by index n, which, according to the calculated estimate, with a possible error of up to ± 5%,
h _p (Q _and ) = [h _p ^nom (Q _and ) ± 0.02] mm,
where the nominal value of h _p ^nom (Q _and ) = 0.06 mm,
which was obtained by pouring according to the scheme: at the inlet, a throttle packet with the value of the initial spill Q _and = 200 cm ³ / min under a pressure drop of 1.0 MPa, spilled at the input overpressure (1 ± 0.02) MPa, at the outlet, spilled at the input overpressure [(1 ± 0.02) - (0.05 ± 0.01)] MPa, and the mechanical term marked by index m, having a value of
h _{input m} = (h _{input m} ^nom ± 0.02) mm
at the input gap of the binary amplifier, which controls the pressure in the chamber of the piston pump drive from the direction of the force of the spring assembly (in the spring chamber), where
h _{input m} ^nom > {0.05+ [h _p ^nom (Q _and ) -h _p ^nom (Q)]} mm,
at the output gap of the binary amplifier, which controls the pressure in the spring chamber, adopted as the base for pouring settings,
h _out.m = (h _out.m ^nom ± 0.00) mm,
where is the face value
h _vyh.m 0,00+ ^nom = {[h _f ^nom (Q _and) -h _f ^nom (Q)]} mm
at the input gap of the binary amplifier, which controls the pressure in the chamber of the piston pump drive from the side of the transmitting rod (in the rod chamber)
h _{input m} ¹ = (h _input ^{m 1nom} ± 0.02) mm,
where is the face value
h _vh.m ^1nom 0,00+ = {[h _f ^nom (Q _and) -h _f ^nom (Q)]} mm
at the output gap of the binary amplifier controlling the pressure in the rod chamber
h _out.m ¹ = h _out.m ^1nom ± 0.02) mm,
where is the face value
h _vyh.m ^1nom> {0,05+ [h _f ^nom (Q _and) -h _f ^nom (Q)]} mm
moreover, the actual value of the spill Q of the throttle package is selected according to the inequality Q≤Q _and , along with this, the spring assembly is made of one spring with a stiffness value of C _p = 4.4 ^+0.5 N / mm. 2. A system for automatically controlling the frequency of rotation of a cascade of a gas turbine engine, comprising an electronic control part and an electronic-hydromechanical actuator, characterized in that the actuator includes a sequentially electronic-hydromechanical drive circuit of the fuel metering device (dispenser) and a hydromechanical power-metering circuit, containing in series the metering portion of the metering unit with a metering hole having an area

where index and is the initial, spool valve for regulating the differential pressure at the metering metering hole (differential valve), the piston drive of the inclined washer of the feed axial-plunger pump (piston pump drive) with the effective piston area

controlled by a differential valve and equipped with a spring assembly with a force directed along the axis of the piston and the transfer rod towards the inclined washer, a pump connected by a power channel to the fuel tanks through a motor filter, while the master part of the differential valve is designed to set the differential pressure

the command part of the differential valve is made of four pairs of radially opposite longitudinal sharp-edged grooves on the valve sleeve with an acceptable connection of the intermediate grooves with the total width of each pair of grooves having the value e = (e ^nom ± 0.1) mm, where the nominal value e ^nom ≥0.6 mm, of two sharp-edged grooves on the spool and is included in the power channel parallel to the metering part of the dispenser, due to which, by means of pouring control and setting the axial position of the four ends of the grooves and mechanical refinement of three of them relative to during operation of the groove edges of the grooves, automatic installation of a rigid connection of two binary hydraulic amplifiers with throttle openings in the form of open gaps (with gaps) is provided, the current values of the width of which are respectively determined by the values of the width h of the installation production gaps, which consist of the index term given by the index and having, according to the calculated estimate with a possible error of up to ± 5% value
h _p (Q _and ) = [h _p ^nom (Q _and ) ± 0.02] mm,
where the nominal value of h _p ^nom (Q _and ) = 0.06 mm,
which is obtained by pouring according to the scheme: at the inlet of the throttle package with the value of the initial spill Q _and = 200 cm ³ / min under a pressure drop of 1.0 MPa, spilled at the input overpressure (1 ± 0.02) MPa, at the outlet there is a gap, spilled at the input overpressure [(1 ± 0.02) - (0.05 ± 0.01)] MPa, and the mechanical term marked by index m, which has a value at the input gap of the binary amplifier controlling the pressure in the chamber of the piston pump drive from the direction of the force of the spring unit (in the spring chamber) -
h _{input m} = (h _{input m} ^nom ± 0.02) mm,
where is the face value
h _{input m} ^nom > {0.05+ [h _p ^nom (Q _and ) - h _p ^nom (Q)]} mm,
at the output gap of the binary amplifier, which controls the pressure in the spring chamber, adopted as the base for pouring settings
h _out.m = (h _out.m ^nom ± 0.00) mm,
where is the face value
h _vyh.m 0,00+ ^nom = {[h _f ^nom (Q _and) - h _f ^nom (Q)]} mm
at the input gap of the binary amplifier that controls the pressure in the chamber of the piston pump drive from the side of the transmitting rod (in the stock chamber) -
h _{input m} ¹ = (h _input ^{m 1nom} ± 0.02) mm,
where is the face value
h _vh.m ^1nom 0,00+ = {[h _f ^nom (Q _and) - h _f ^nom (Q)]} mm
at the output gap of the amplifier controlling the pressure in the rod chamber -
h _out.m ¹ = (h _out.m ^1nom ± 0.02) mm,
where is the face value
h _vyh.m ^1nom> {0,05+ [h _f ^nom (Q _and) - h _f ^nom (Q)]} mm
moreover, the actual value of the spill Q of the throttle package is selected according to the inequality Q≤Q _and , along with this, the spring assembly is made from the number of springs installed in parallel from row 1, ... n and the number of series installed from row 0.2, ... m springs and at the same time equipped with, respectively, the number of the latter stops corresponding to the values of the positions of the piston of the piston drive of the pump (piston) Н ₂ , ..., Н _m , which provides the node with the current value of effective stiffness

where H is the current position of the piston;
H _{j + 1} - (j + 1) -th position of the piston;

rigidity of the i-th parallel connected spring;

stiffness of the (j + 1) th spring connected in series,

the stiffness of the first spring connected in series,
at

at

exceeding a value of 4.9 N / mm, with an optimal range of values determined by the inequality of the sum of the force gradients

where N _H (H × n _inH ) is the force from the moment on the inclined pump washer acting along the axis of the piston pump drive, the mode values of which are determined experimentally, where n _inH is the speed of the high pressure cascade reduced to the pump drive;
N _P (N) is the force created by the spring unit, as a result of which ∂N _P (H) / ∂H = C _P (H),
in connection with which the actual values of the values of s _d , S _by , P _Δ are selected on the basis of relationships that are subject to the following conditions:

or

where N _N (N) is the force from the moment on the inclined washer of the pump at n _vDH , providing a maximum of the sum of the forces [N _N (N) + N _P (N)],
provided that in the range of values of s _d the equality

3. A system for automatically controlling the speed of a cascade of a gas turbine engine, comprising a control part and an electronic-hydromechanical actuator, characterized in that the actuator includes a series of electronic-hydromechanical drive circuit fuel dispenser (dispenser) and a hydromechanical power-metering circuit containing sequentially dispensing part of the dispenser, spool valve for regulating the differential pressure at the dispensing hole of the dispenser (differential valve), pore non-rotary drive of the inclined washer of the feed axial-plunger pump (piston drive of the pump), controlled by the differential valve and equipped with a spring assembly with a force directed along the axis of the piston and the transfer rod towards the inclined washer, the pump connected to the fuel supply channel through the engine filter by the driver of the differential valve is designed to set the differential pressure, the command part of the differential valve is made of two pairs of radially opposite longitudinal sharp-edged grooves on the valve sleeve with a total the width of each pair of grooves e = (e ^nom ± 0.1) mm, where the nominal e ^nom ≥0.6 mm, one sharp-edged groove on the spool and is included in the power channel parallel to the metering portion of the dispenser so that it is connected to the fuel supply channel in the first sub-option or the fuel exhaust channel in the second sub-variant and, accordingly, with the fuel drain through the channels with constant chokes switched on with a spill Q _s , where the index c is constant, in the first sub-variant or with the fuel supply through channels with constant chokes switched on with a Q _s spill in the second podvar Iante, due to which, through pouring control and setting the axial position of both ends of the grooves and mechanical refinement of one of them, automatic operation of a rigid connection of two single hydraulic amplifiers with working holes of variable chokes in the form of open gaps (with gaps), the current width values of which are determined, is ensured during operation the values of the width h of the installation production gaps, consisting of the pouring term indicated by the index n, having, according to the estimated with a possible error of up to ± 5%, value
h _p (Q _and ) = [h _p ^nom (Q _and ) ± 0.02] mm,
where the nominal value of h _p ^nom (Q _and ) = 0.06 mm,
which was obtained by pouring according to the scheme: at the inlet, a throttle packet with the value of the initial spill Q _and = 200 cm ³ / min under a pressure drop of 1.0 MPa, spilled at the input overpressure (1 ± 0.02) MPa, at the outlet, a gap, spilled at the input overpressure [(1 ± 0.02) - (0.05 ± 0.01)] MPa, and the mechanical term marked by index m, having a value of one of the gaps of the variable chokes
h _m1 = (h _m1 ^nom ± 0.00) mm,
where is the face value
^SG h = _{m1 0,00+ [h _f ^nom (Q _and) -h _f ^nom (Q)]} mm
another
h _m2 = (h _m2 ^nom +0.02) mm,
where is the face value
h _m2 ^nom ≥ {0.02+ [h _n ^nom (Q _and) - h _n ^nom (Q)]} mm,
the value of the spill of constant chokes is selected under a pressure drop of 1.0 MPa according to the equality

where e ^nom in mm
moreover, the actual value of the spill Q of the throttle package is selected according to the inequality Q ≤ Q _and ,
along with this, the spring assembly is made of one spring with a stiffness value of C _p = 4.4 ^+0.5 N / mm. 4. A system for automatically controlling the frequency of rotation of a cascade of a gas turbine engine, comprising an electronic control part and an electronic-hydromechanical actuating part, characterized in that the executive part includes a sequentially electronic-hydromechanical low-inertia drive circuit of the fuel flow metering device (dispenser) and a hydromechanical feed-metering circuit containing sequentially dosing part of the dispenser with a dosing hole with an area

where index and is the initial, spool valve for regulating the differential pressure at the metering metering hole (differential valve), the piston drive of the inclined washer of the feed axial-plunger pump (piston pump drive) with the effective piston area

controlled by differential valves and equipped with a spring assembly with a force directed along the axis of the piston and the transmission rod towards the inclined washer, a pump connected by a power channel to the fuel tanks through a motor filter, while the differential valve master portion is designed to set the differential pressure

the command part of the differential valve is made of two pairs of radially opposite longitudinal sharp-edged grooves on the valve sleeve with a total width of each pair of grooves e = (e ^nom ± 0.1) mm, where the nominal ^{number is} ≥0.06 mm, one sharp-edged groove on the spool and included in the power channel parallel to the metering portion of the dispenser so that it is connected to the fuel supply channel in the first sub-option or the fuel drain channel in the second sub-option and, accordingly, to the fuel drain through the channels with constant chokes turned on with a flow Q _s , where ind Exc c - constant, in the first sub-variant or with fuel supply through channels with constant throttles switched on with Q _s in the second sub-variant, thanks to which automatic installation of rigid installation is ensured during operation by pouring control and setting the axial position of both ends of the groove and mechanical refining of one of them bundles of two single hydraulic amplifiers with working openings of variable chokes in the form of open gaps (with gaps), the current values of the width of which are determined by the values of w irins h of adjusting production gaps, consisting of the pouring term marked by index n, which, according to the estimated estimate, with a possible error of up to ± 5%,
h _p (Q _and ) = [h _p ^nom (Q _and ) ± 0.02] mm,
where is the face value
h _p ^nom (Q _and ) = 0.06 mm,
which is obtained by pouring according to the scheme: at the inlet of the throttle package with the value of the initial spill Q _and = 200 cm ³ / min under a pressure drop of 101 MPa, spilled at the input overpressure (1 ± 0.02) MPa, at the outlet, the gap spilled at inlet overpressure [(1 ± 0.02) - (0.05 ± 0.01)] MPa, and the mechanical term marked by index m, having a value equal to one of the gaps of the measured chokes
h _m1 = (h _m1 ^nom ± 0.00) mm,
where is the face value
^SG h = _{m1 0,00+ [h _f ^nom (Q _and) - h _f ^nom (Q)]} mm
another
h _m2 = (h _m2 ^nom ± 0.02) mm,
where is the face value
h _m2 ^nom ≥ {0.02+ [h _n ^nom (Q _and ) - h _n ^nom (Q)]} mm,
the value of the spill of constant chokes is selected under a pressure drop of 1.0 MPa according to the equality

where e ^nom in mm
moreover, the actual value of the spill Q of the technological throttle package is selected according to the inequality Q ≤ Q _and ,
along with this, the spring assembly is made up of a number of springs installed in parallel from a row 1, ... n and a number of springs installed in a row from 0.2, ... m selected from a row, and at the same time is equipped with a number of last stops corresponding to piston piston positions pump (piston) drive Н ₂ , ..., Н _m , which provides the node with the current value of effective stiffness

where H is the current position of the piston;
H _{j + 1} - (j + 1) -th position of the piston;

rigidity of the i-th parallel connected spring;

stiffness of the (j + 1) th spring connected in series;

the stiffness of the first spring connected in series,
at

at

exceeding a value of 4.9 N / mm, with an optimal range of values determined by the inequality of the sum of the force gradients

where N _H (H × n _inH ) is the force from the moment on the inclined pump washer acting along the axis of the piston drive, the mode values of which are determined experimentally, where n _inH is the speed of the high pressure cascade reduced to the pump drive;
N _P (N) is the force created by the spring unit, as a result of which ∂N _P (H) / ∂H = C _P (H),
in connection with which the actual values of the values of s _d , S _by , P _Δ are selected on the basis of relationships that are subject to the following conditions:

or

where N _N (N) is the force from the moment on the inclined washer of the pump at n _vDH , providing the maximum sum of forces [N _P (N) + N _P (N)],
provided that in the range of values of s _d the equality

5. A system for automatically controlling the speed of a cascade of a gas turbine engine, comprising an electronic control part and an electronic-hydromechanical actuator, characterized in that the actuator includes a series of electronic-hydromechanical drive circuit fuel dispenser (dispenser) and a hydromechanical feed-metering circuit containing sequentially metering part of the metering valve, differential pressure control valve on the metering metering hole (valve turnip), piston drive inclined washers of the feed axial plunger pump (piston pump drive), controlled by the differential valve and equipped with a spring assembly with a force directed along the axis of the piston and the transmission rod towards the inclined washer, a pump connected to the fuel supply channel through the engine a filter, while the differential valve master portion is designed to set the differential pressure, the differential valve command part is made of two pairs of radially opposite longitudinal sharp-edged grooves on the valve sleeve ana with a total width of each pair of grooves e = (f ^nom ± 0,1) mm, where f ^nom denomination ≥0,6 mm, two sharp-edged grooves on the spool and is included in the feed channel in parallel with a dispensing portion of the dispenser providing, along with it, or connecting the chamber of the piston pump drive from the direction of the force of the spring assembly with a channel with the outlet from the metering opening of the dispenser in the first variant or connecting the chamber of the piston drive of the pump from the side of the piston drive rod with a channel with the entrance to the metering hole of the meter in the second variant e, due to which, by pouring control and setting the axial position of two symmetrical ends of the grooves and mechanical refinement of one of them relative to the side edge of the corresponding groove, automatic installation of one binary hydraulic amplifier with throttle holes in the form of open gaps (with gaps): or gaps, is ensured during operation adjacent to the ends of the jumper between the grooves in the first sub-option, or gaps adjacent to the extreme ends of the grooves in the second sub-option, the current eniya width which define widths h industrial installation clearances consisting of the marked index n prolivochnogo term having calculated according to a possible estimation error of + 5%, the value of
h _p (Q _and ) = [h _p ^nom (Q _and ) ± 0.02] mm,
where the nominal value of h _p ^nom (Q _and ) = 0.06 mm,
which is obtained by pouring according to the scheme: at the inlet, a throttle package with the value of the initial spill Q _and = 200 cm ³ / min under a pressure drop of 1.0 MPa, spilled at the input overpressure (1 ± 0.02) MPa, at the outlet, spilled at the input overpressure [(1 ± 0.02) - (0.05 ± 0.01)] MPa, and the mechanical term marked by index m, which has a value of one of the gaps, for example, the input, as the base for pouring tincture
h _{input m} = (h _{input m} ^nom ± 0.00) mm,
where is the face value
h _vh.m ^prefecture> {0,00+ [h _f ^nom (Q _and) -h _f ^nom (Q)]} mm
another
h _out.m = (h _out.m ^nom ± 0.02) mm,
where is the face value
h _{output m} ^nom > {0.03+ [h _n ^nom (Q _and ) -h _n ^nom (Q)]} mm,
moreover, the actual value of the spill Q of the throttle package is selected according to the inequality Q≤Q _and along with this, the spring assembly is made of one spring with the value of effective stiffness C _P = 4.4 ^+0.5 N / mm. 6. A system for automatically controlling the speed of a cascade of a gas turbine engine, comprising an electronic control part and an electronic-hydromechanical actuator, characterized in that the actuator includes a series of electronic-hydromechanical drive circuit fuel dispenser (dispenser) and a hydromechanical power-metering circuit containing sequentially the metering portion of the metering unit with a metering hole having an area

where index and is the initial, spool valve for regulating the differential pressure at the metering metering hole (differential valve), the piston drive of the inclined washer of the feed axial-plunger pump (piston pump drive) with the effective piston area

controlled by a differential valve and equipped with a spring assembly with a force directed along the axis of the piston and the transfer rod towards the inclined washer, a pump connected by a power channel to the fuel tanks through a motor filter, while the master part of the differential valve is designed to set the differential pressure

the command part of the differential valve is made of two pairs of radially opposite longitudinal sharp-edged grooves on the valve sleeve with the total width of each pair of grooves e = (e ^nom ± 0.1) mm, where the nominal ^{number is} ≥0.6 mm, two sharp-edged grooves on the spool and included in the power channel parallel to the metering portion of the dispenser, while providing, at the same time, either connecting the piston drive chamber of the pump from the direction of the force of the spring assembly to the channel with the outlet from the metering dispensing hole in the first variant or connecting the piston chamber of the pump drive from the piston drive rod side with a channel with an entrance to the dispensing hole of the dispenser in the second variant, due to which, through operation, the axial position of two symmetrical ends of the grooves and mechanical refinement of one of them relative to the side edge of the groove are ensured, automatic operation of one binary hydraulic an amplifier with throttle openings in the form of open gaps (with gaps): or gaps adjacent to the ends of the jumper between the groove in the first embodiment, or gaps adjacent to the extreme ends of the bores in a second sub-embodiment, the current widths of which are defined widths h industrial installation clearances consisting of the marked index n prolivochnogo term having calculated according to a possible estimation error of ± 5% significance
h _p (Q _and ) = [h _p ^nom (Q _and ) ± 0.02] mm,
where the nominal value of h _p ^nom (Q _and ) = 0.06 mm,
which was obtained by pouring according to the scheme: at the inlet, a throttle packet with the value of the initial spill Q _and = 200 cm ³ / min under a pressure drop of 1.0 MPa, spilled at the input overpressure (1 ± 0.02) MPa, at the outlet, a gap, spilled at the input overpressure [(1 ± 0.02) - (0.05 ± 0.01)] MPa, and marked by the index m of the mechanical term, which has a value of one of the gaps, for example, the input, as the base for the pouring setting ,
h _{input m} = (h _{input m} ^nom ± 0.00) mm,
where is the face value
h _{input m} ^nom > {0.01+ [h _n ^nom (Q _and ) -h _n ^nom (Q)]} mm,
another
h _out.m = (h _out.m ^nom ± 0.02) mm,
where is the face value
h _{output m} ^nom ≥ {0.03+ [h _p ^nom (Q _and ) -h _p ^nom (Q)]} mm,
moreover, the actual value of the spill Q of the throttle package is selected according to the inequality Q≤Q _and ,
along with this, the spring assembly is made of a number of springs installed in parallel from a row 1, ... n and a number of springs installed in a row from 0.2, ... m selected from a row, and at the same time it is equipped with a number of latter stops corresponding to the values of the positions of the piston actuator pump (piston) H ₂ , ..., H _m , which provides the node with the current value of effective stiffness

where H is the current position of the piston;
H _{j + 1} - (j + 1) -e - the position of the piston;

rigidity of the i-th parallel connected spring;

stiffness of the (j + 1) th spring connected in series;

the stiffness of the first spring connected in series,
at

at

exceeding a value of 4.9 N / mm, with an optimal range of values determined by the inequality of the values of the sum of the force gradients

where N _H (H × n _inH ) is the force from the moment on the inclined pump washer acting along the axis of the piston pump drive, the mode values of which are determined experimentally, where n _inH is the speed of the high pressure cascade reduced to the pump drive;
N _P (N) is the force created by the spring unit, as a result of which ∂N _P (H) / ∂H = C _P (H),
in connection with which the actual values of the values of s _d , S _by , P _Δ are selected on the basis of relationships that are subject to the following conditions:

or

where N _H (H) is the force from the moment on the inclined washer of the pump at n _vDH , providing the maximum sum of forces [N _H (H) + N _P (N)],
provided that in the range of values of s _d the equality

7. A system for automatically controlling the speed of a cascade of a gas turbine engine, comprising an electronic control part and an electronic-hydromechanical actuator, characterized in that the actuator includes a series of electronic-hydromechanical drive circuit fuel dispenser (dispenser) and a hydromechanical feed-metering circuit containing sequentially dosing part of the dispenser, spool valve for regulating the differential pressure at the dispensing opening of the dispenser (valve turnip), a piston drive of an inclined washer of a feed axial plunger pump (piston drive of a pump), controlled by a differential valve and equipped with a spring assembly with a force directed along the axis of the piston and a transmission rod towards the inclined washer, a pump connected to the fuel supply channel through the engine a filter, while the differential valve master portion is designed to set the differential pressure, the differential valve command part is made of one pair of radially opposite longitudinal sharp-edged grooves on the sleeve Pan with a total width e = (f ^nom ± 0,1) mm, where f ^nom denomination ≥0,6 mm, a sharp-edged groove on the spool and is included in the supply channel parallel to the dispensing portion of the dispenser so that the fuel supply is connected to the first channel and second sub-options or with a fuel exhaust channel in the third and fourth sub-options and, respectively, with a fuel exhaust channel with a constant throttle in the first and second sub-options or with a fuel feed channel with a constant throttle in the third and fourth sub-options to provide a connection the chamber of the piston drive of the pump from the direction of the force of the spring assembly with a channel exiting the metering opening of the dispenser in the first and third variants or the connection of the chamber of the piston drive of the pump from the side of the transmitting rod with the entrance to the metering hole of the meter in the second and fourth variants, which ensures operation automatic installation of one single hydraulic amplifier with a variable choke in the form of an open gap, the current values of the width of which are determined by the value Q spilling it _with a fixed orifice pressure drop of 1.0 MPa, selected according to inequality

where e ^nom in mm
along with this, the spring assembly is made of one spring with a value of effective stiffness C _P = 4.4 ^+0.5 N / mm. 8. A system for automatically controlling the speed of a cascade of a gas turbine engine, comprising an electronic control part and an electronic-hydromechanical actuator, characterized in that the actuator includes a series of electronic-hydromechanical drive circuit fuel dispenser (dispenser) and a hydromechanical power-metering circuit containing sequentially the metering portion of the metering unit with a metering hole having an area

where index and is the initial, spool valve for regulating the differential pressure at the metering metering hole of the metering device, the piston drive of the inclined washer of the feed axial-plunger pump (piston pump drive) with the effective piston area

controlled by a differential valve and equipped with a spring assembly with a force directed along the axis of the piston and the transfer rod towards the inclined washer, a pump connected by a power channel to the fuel tanks through a motor filter, while the master part of the differential valve is designed to set the differential pressure

the command part of the differential valve is made of one pair of radially opposite longitudinal sharp-edged grooves on the valve sleeve with a total width of e = (e ^nom ± 0.1) mm, where the nominal ^{number is} ≥0.6 mm, one sharp-edged groove on the spool and is included in the channel power supply parallel to the metering part of the dispenser so that it is connected to the fuel supply channel in the first and second sub-options or to the fuel drain channel in the third and fourth sub-options and, accordingly, to the fuel exhaust channel with a constant choke in the first and second supply options or with a fuel supply channel with a constant throttle in the third and fourth sub-options, providing a connection of the piston drive chamber of the pump from the direction of the force of the spring assembly with a channel exit from the metering opening of the dispenser in the first and third options or connecting the piston drive chamber of the pump on the side of the transmitting rod with the entrance to the dispensing hole of the dispenser in the second and fourth sub-options, which ensures the automatic installation of one single g dravlicheskogo amplifier with a variable throttle in the form of an open gap, the width of the current values of which are determined spilling Q value _with a fixed orifice pressure drop of 1.0 MPa, selected according to inequality

where e ^nom in mm
at the same time, the spring assembly is made of a number of springs installed in parallel from a row 1, ..., n and a number of springs installed in series from a row of 0.2, ..., m, and at the same time is equipped with a number of latter stops corresponding to position the piston of the piston drive of the pump (piston) H ₂ , ..., H _m , which provides the node with the current value of effective stiffness

where H is the current position of the piston;
H _{j + 1} - (j + 1) -e - the position of the piston;

rigidity of the i-th parallel connected spring;

stiffness of the (j + 1) th spring connected in series;

the stiffness of the first spring connected in series,
at

at

exceeding a value of 4.9 N / mm, with an optimal range of values determined by the inequality of the sum of the force gradients

where N _H (H × n _inH ) is the force from the moment on the inclined pump washer acting along the axis of the piston pump drive, the mode values of which are determined experimentally, where n _inH is the speed of the high pressure cascade reduced to the pump drive;
N _P (N) is the force created by the spring unit, as a result of which ∂N _P (H) / ∂H = C _P (H),
in connection with which the actual values of the values of s _d , S _by , P _Δ are selected on the basis of relationships that are subject to the following conditions

or

where N _N (N) is the force from the moment on the inclined washer of the pump at n _vDH , providing the maximum sum of forces [N _H (H) + N _P (H)],
provided that in the range of values of s _d the equality

u

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