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    • 专利摘要:
    A pump system includes a pressure compensator and a speed compensator for controlling the displacement of a variable displacement pump having an inlet and an outlet. The variable displacement pump is driven by a drive shaft powered by an electric motor. The pressure compensator adjusts the displacement of the variable displacement pump according to a pump pressure at the pump outlet. The speed compensator adjusts a maximum amplitude of the displacement of the variable displacement pump according to an electric motor speed.
    公开号:FR3075278A1
    申请号:FR1871504
    申请日:2018-11-13
    公开日:2019-06-21
    发明作者:Phillip Wayne Galloway
    申请人:Eaton Intelligent Power Ltd；
    IPC主号:
    专利说明:

    Title of the invention: HYDRAULIC PUMP REGULATION Technical field [0001] The present invention relates generally to hydraulic pumps. The present invention relates more particularly to hydraulic pump systems and methods for regulating the torque demand.
    Background [0002] Hydraulic pumps can provide a source of power for flight controls, landing gear and other components of an aircraft. An aircraft generally supplies variable frequency electrical power to hydraulic pumps driven by an electric motor, as the engine speed of an aircraft varies during flight. For example, the engine speed of an aircraft is high during takeoff, and low during landing. Variable frequency electric power can cause problems for electric motors because, as the frequency and the engine speed increase, the motor torque decreases. As a result, electric motors may not have enough torque to drive hydraulic pumps operating at maximum displacement.
    One solution to this problem has been to convert the variable frequency into direct current, then to convert the direct current into an effective alternating current which can be used to power the electric motors which drive the hydraulic pumps. However, complex and cumbersome electronic systems are required for these conversions. In addition, these systems are often located in regions of the aircraft exposed to a high temperature and the heating of the electronics is therefore a major problem.
    Summary [0004] Aspects of the present invention relate to the reduction of a maximum admissible stroke length of a variable displacement hydraulic pump when a predetermined speed threshold of an electric motor is reached. By limiting the maximum admissible stroke length to the speed threshold, a traditional electric motor can be used to drive the variable displacement hydraulic pump.
    In one aspect, the technology described relates to a pump system comprising a variable displacement pump including an inlet and an outlet. The variable displacement pump includes a drive shaft and an electric motor to drive a rotation of the drive shaft. A pressure compensator adjusts a pump displacement of the variable displacement pump according to a pump pressure at the pump outlet, and a speed compensator adjusts a maximum amplitude of the pump displacement inversely proportional to a rotation speed of the drive shaft once the speed of rotation of the drive shaft has exceeded a predetermined speed threshold.
    The pressure compensator may include a pressure compensation actuator adapted to receive hydraulic pressure from the pump outlet when a pressure compensator valve is opened. The pressure compensation actuator deactivates a cylinder head supporting a swash plate in the variable displacement pump when the pump pressure exceeds a predetermined pressure.
    The speed compensator may include a displacement limiting actuator adapted to receive hydraulic pressure from the pump outlet when a speed compensation valve is opened. The speed compensation valve can be a distributor having a variable opening proportional to the speed of the electric motor. The speed compensator includes a fixed displacement pump coupled to the drive shaft and adapted to draw hydraulic fluid from the inlet of the variable displacement pump and to pump hydraulic fluid through an orifice. A pressure differential across the port is proportional to the speed of the electric motor and, when the pressure differential across the port exceeds a predetermined value, a pressure located on an upstream side of the port opens the compensation valve The hydraulic fluid circulates through the orifice in a return line for the cooling of the components of the pump system.
    The variable displacement pump can include a rotary group coupled to the drive shaft so that the rotary group rotates together with the drive shaft about a drive shaft axis. The rotary group includes a piston block defining a plurality of cylinder bores positioned around the drive shaft axis. The rotary unit includes pistons that reciprocate in the cylinder bores to pump hydraulic fluid through the variable displacement pump from inlet to outlet, the pistons having piston shoes.
    The variable displacement pump may further include a cylinder head supporting a swash plate, and the piston shoes of the rotary group pistons are configured to slide along the swash plate while the rotary group rotates around the shaft tax. drive. The cylinder head can pivot to move relative to a drive shaft axis to adjust a swash plate angle of the swash plate to the drive shaft axis.
    The speed compensator can include a displacement limiting actuator designed to move a maximum displacement position of the cylinder head in order to reduce a maximum admissible stroke length authorized by the swash plate, and the [0011] [0012] [ Pressure compensator includes a pressure compensating actuator designed to defuse the cylinder head when a pump pressure at the outlet of the variable displacement pump exceeds a predetermined pressure so that the stroke length of the group pistons rotary is less than the maximum admissible stroke length authorized by the swash plate.
    The pump system may be provided with a negative cylinder head actuator which applies a negative moment force to compensate for a positive moment force on the cylinder head. The electric motor can be designed to receive variable frequency electric energy from a power source on an aircraft.
    In another aspect, the technology described relates to a method for controlling a pump displacement of a variable displacement pump, the variable displacement pump being provided with a cylinder head supporting a swash plate, the cylinder head being driven by an electric motor for rotate around a drive shaft axis and the cylinder head can move pivotally relative to the drive shaft axis. The method may include establishing a maximum cylinder displacement position; determining whether the electric motor speed exceeds a predetermined speed; when the electric motor speed exceeds the predetermined speed, the actuation of a displacement limitation actuator to pivot the cylinder head towards a minimum or zero displacement position to establish a new maximum displacement position which reduces a maximum stroke length admissible authorized by the swash plate; determining whether a pump pressure at a pump outlet of the variable displacement pump exceeds a predetermined pressure; and when the pump pressure exceeds the predetermined pressure, defusing the cylinder head towards the minimum or zero displacement position so that the stroke length allowed by the swash plate is less than the maximum allowable stroke length.
    The step of determining whether the electric motor speed exceeds the predetermined speed can be performed by determining whether a differential pressure through an orifice exceeds a predetermined value. The step of actuating the displacement limitation actuator can be accomplished by opening a speed compensation valve to apply hydraulic pressure from the pump output of the variable displacement pump to the displacement limitation actuator.
    In some examples, the size of the opening of the speed compensation valve varies according to the amplitude of the differential pressure through the orifice which is proportional to the speed of the electric motor. In some examples, the method includes applying a negative moment force to the cylinder head to compensate for a positive moment force on the cylinder head resulting from an increase in the speed at which the cylinder head rotates about the axis of drive shaft.
    In another aspect, the technology described relates to a pump system comprising a variable displacement pump which includes a pump housing defining a pump inlet and a pump outlet. A drive shaft extends through the pump housing, the drive shaft being rotatable about a drive shaft axis. A rotary group inside the pump housing is coupled to the drive shaft so that the rotary group rotates together with the drive shaft around the drive shaft axis. The rotary group includes a piston block defining a plurality of cylinder bores positioned around the drive shaft axis. The rotary unit also includes pistons with piston shoes which reciprocate in the cylinder bores to pump hydraulic fluid through the in-line pump from the pump inlet to the pump outlet. The pistons are fitted with piston shoes.
    The variable pump system further includes a cylinder head inside the pump housing which supports a swash plate, and the piston shoes of the rotary group pistons are configured to slide along the swash plate while the group rotary rotates around the drive shaft axis. The cylinder head can be pivotally moved relative to the drive shaft axis to adjust a swash plate angle of the swash plate relative to the drive shaft tax. The swash plate angle establishes a displacement of the variable displacement pump by rotation of the rotary unit by establishing a stroke length of the pistons in the cylinder bores. The cylinder head can move pivotally with respect to the drive shaft axis between a maximum displacement position and a minimum or zero displacement position. The pistons have a maximum stroke length when the cylinder head is in the maximum displacement position and a minimum stroke length or no stroke length when the cylinder head is in the minimum or zero displacement position. The cylinder head is biased by a spring towards the maximum displacement position.
    A pressure compensator adjusts the pump displacement according to a pump pressure at the pump outlet. The pressure compensator includes a pressure compensation actuator for moving the cylinder head in a defusing direction from the maximum displacement position to the minimum or zero displacement position. The pressure compensator also includes a pressure compensating line for applying hydraulic pressure to the pressure compensating actuator to move the cylinder head in the defusing direction. The pressure compensator also includes a pressure compensator valve for opening and closing a fluid communication between the pressure compensating line and the pump outlet. The pressure compensator valve can move between a closed position in which a fluid communication [0018] [0020] [0021] [0022] [0023] is blocked between the pump outlet and the pressure compensation line , and an open position in which a fluid communication is open between the pump outlet and the pressure compensation line, The pressure compensator valve is biased to the closed position by a biasing force. The pressure compensator valve is configured to open when the pump pressure at the pump outlet exceeds the predetermined pressure.
    The pump system may further include an electric motor to drive a rotation of the drive shaft, a speed compensator which includes an orifice and a fixed displacement pump coupled to the drive shaft so that the drive shaft rotates the fixed displacement pump. The fixed displacement pump is configured to draw hydraulic fluid from an inlet side of the variable displacement pump and to pump hydraulic fluid through the port. A speed compensation line applies hydraulic pressure to the cylinder limiting actuator to move the cylinder head maximum cylinder position to the minimum or zero cylinder position, so that a new maximum cylinder position is established and that a maximum admissible stroke length authorized by the swash plate is reduced. A speed compensation valve opens and closes fluid communication between the pump outlet and the speed compensation line. The speed compensation valve is biased to a closed position where fluid communication between the pump outlet and the speed compensation line is blocked, and the speed compensation valve is configured to open fluid communication between the pump outlet and the speed compensation line when a differential pressure through the orifice exceeds a predetermined value.
    Various additional aspects of the invention will be set out in the description below. Aspects of the invention may relate to individual features and a combination of features. It should be understood that the general description above and the detailed description below are given by way of example and explanation only and do not limit the general concepts of the invention on which the embodiments described here are based. .
    Brief description of the drawings
    FIG. 1 is a schematic diagram of a hydraulic fluid circuit of an aircraft.
    Figure 2 is a schematic diagram of components downstream of the aircraft.
    FIG. 3 is a schematic diagram of fluid circuits of the aircraft.
    Figure 4 is a schematic diagram of a hydraulic pump system in operation at low speed.
    [0025] [0027] [0028] [0028] [0029] [0030] [0031] [0032] [0033] [0034] [0035] [0036] [0037] [0038] [0039]
    Figure 5 is a schematic diagram of the hydraulic pump system in high speed operation.
    FIG. 6 represents a displacement limitation actuator and a pressure compensation actuator in operation, at low speed and at a maximum admissible stroke length.
    FIG. 7 represents the displacement limitation actuator and the pressure compensation actuator in operation at low speed and at an intermediate stroke length.
    FIG. 8 represents the displacement limitation actuator and the pressure compensation actuator in operation at medium speed and at a maximum stroke length.
    FIG. 9 represents the displacement limitation actuator and the pressure compensation actuator in operation at medium speed and at an intermediate stroke length.
    FIG. 10 represents the displacement limitation actuator and the pressure compensation actuator in operation at high speed and at a maximum stroke length.
    FIG. 11 shows the displacement limitation actuator and the pressure compensation actuator in operation at high speed and at an intermediate stroke length.
    FIG. 12 is a graph illustrating a maximum admissible stroke with respect to an electric motor speed.
    FIG. 13 represents a negative cylinder head actuator in operation at low speed and at a maximum admissible stroke length.
    FIG. 14 represents the negative cylinder head moment actuator in operation at low speed and at an intermediate stroke length.
    FIG. 15 shows the negative cylinder head actuator in operation at medium speed and at a maximum stroke length.
    Figure 16 shows the negative cylinder head actuator when operating at medium speed and an intermediate stroke length.
    Figure 17 shows the negative cylinder head actuator in operation at high speed and at maximum stroke length.
    FIG. 18 shows the negative cylinder head actuator in operation at high speed and at an intermediate stroke length.
    Figure 19 is a graph illustrating the cylinder head moment with respect to the electric motor speed.
    Figure 20 is a block diagram illustrating a method of controlling the displacement of a variable displacement pump.
    Various embodiments will be described in detail with reference to the drawings, in which identical reference numerals represent identical parts and assemblies in the various views. The reference to various embodiments does not limit the scope of the appended claims. In addition, none of the examples presented in this description are intended to be limiting and these examples simply present some of the many possible embodiments for the appended claims.
    Figure 1 is a schematic diagram of an aircraft 13 having a hydraulic fluid circuit 10. As shown in Figure 1, the hydraulic fluid circuit 10 is located in a fuselage 11 of the aircraft 13. The aircraft 13 can include multiple or redundant fluid circuits. The fluid circuit 10 includes at least one hydraulic pump system 12 and a cooling circuit 14 which is in fluid communication with the hydraulic pump system 12.
    Also located in the fuselage 11 of the aircraft 13, an energy source 20 drives the hydraulic pump system 12. The energy source 20 can be an aircraft engine or an engine separate from the engine. aircraft. In some examples, the energy supplied by the energy source 20 to the hydraulic pump system 12 is variable frequency electrical energy which can vary during the flight of the aircraft 13.
    The hydraulic pump system 12 uses hydraulic fluid obtained from a fluid reservoir 18 to drive active components downstream 16 (for example, actuators, cylinders, steering units, motors, valves, etc.) of the aircraft
    13. These downstream components 16 can be used to power the flight controls, the landing gear and other components of the aircraft 13. Although the hydraulic pump system 12 is described herein as being suitable for use in aircraft applications, it will be understood that the hydraulic pump system 12 can be used for other applications as well.
    Figure 2 is a schematic diagram of the aircraft 13 having downstream components 16 connected to multiple fluid circuits. As shown in FIG. 2, each downstream component 16 is connected to one or more fluid circuits, such as the first and second fluid circuits X, Z given by way of example, so that, in the case where a fluid circuit breaks down or may work, a backup or redundant fluid circuit can be used to supply the downstream component 16. Although the downstream components 16 are represented on the wings, the elevators and the control surface direction of the aircraft 13, it will be understood that the downstream components 16 can also be located on other parts of the aircraft 13.
    Figure 3 is a schematic diagram of fluid circuits X, Z. As shown in Figure 3, each fluid circuit X, Z can include one or more hydraulic pump systems 12. For example, the fluid circuit X can have three hydraulic pump systems 12 and the fluid circuit Z can have three hydraulic pump systems 12. Multiple hydraulic pump systems 12 add another level of redundancy, so that in the event of failure or malfunction of a hydraulic pump system 12, a backup or redundant hydraulic pump system 12 can be used to supply fluid to the fluid circuits X, Z.
    Figure 4 is a schematic diagram of a hydraulic pump system 12 in low speed operation. As shown in Figure 4, the hydraulic pump system 12 includes a housing 100 which includes a pump inlet 126 and a pump outlet 128 for a variable displacement pump 102. A drive shaft 106 extends through the housing 100. The drive shaft 106 is coupled to an electric motor 146 designed to rotate the drive shaft 106 around a drive shaft axis 122. The electric motor 146 can be powered by the source energy 20 of the aircraft 13 illustrated in FIG. 1.
    A rotary group 104 is coupled to the drive shaft 106. The rotary group 104 includes a piston block 108 which defines a plurality of cylinder bores 110 positioned around the drive shaft axis 122 The piston block 108 is coupled to the drive shaft 106 by a mechanical connection which rotatably fixes the piston block 108 to the drive shaft 106 so that the piston block 108 and the rest of the rotary group 104 rotate together with the drive shaft 106 around the drive shaft axis 122. The pistons 112 provided with piston bodies 132 are designed to perform a reciprocating movement in the cylinder bores 110 for pumping hydraulic fluid through the variable displacement pump 102 from the pump inlet 126 to the pump outlet 128. Each piston 112 includes a piston shoe 114 fixed to one end of the piston body 132 and extending from the pisto block ns 108. The piston shoes 114 slide along a swash plate 116 while the rotary group 104 rotates around the drive shaft axis 122, and the pistons 112 perform a reciprocating movement in their corresponding cylinder bores 110 while a relative rotation takes place between the swash plate 116 and the rotary group 104.
    The swash plate 116 is supported by a yoke 130 which can pivotally move relative to a drive shaft axis 122 in order to adjust a swash plate angle θ of the swash plate 116 relative to the drive shaft axis 122. The swash plate angle θ determines the displacement of the variable displacement pump 102 by rotation of the rotary group 104 by establishing a stroke length of the pistons 112 in the cylinder bores 110. The cylinder head 130 can move pivotally with respect to the drive shaft axis 122 between a position of maximum displacement D o and a position of minimum or zero displacement D i (see FIG. 5) . The cylinder head 130 is biased towards the maximum displacement position D 0 . In some examples, the cylinder head 130 is biased by a cylinder head position return spring 162. The pistons 112 inside the cylinder bores 110 have a maximum stroke length when the cylinder head 130 is in the maximum displacement position D o and the pistons 112 have a minimum stroke length or no stroke length when the cylinder head 130 is in the minimum or zero displacement position Dj. The maximum stroke length causes a maximum displacement of the variable displacement pump 102 by rotation of the rotary unit 104, while the minimum stroke length causes a minimum or zero displacement of the variable displacement pump 102 by rotation of the rotary unit 104.
    Still with reference to FIG. 4, a pressure compensator 134 adjusts the displacement of the variable displacement pump 102 according to a pump pressure at the pump outlet 128. The pressure compensator 134 includes a pressure compensator valve 144 to open and close a fluid communication between a pressure compensation line 142 and the pump outlet 128. The pressure compensator valve 144 can move between a closed position in which a fluid communication is blocked between the pump outlet 128 and the pressure compensation line 142, and an open position in which a fluid communication is open between the pump outlet 128 and the pressure compensation line 142. The pressure compensator valve 144 is biased toward the closed position by a biasing element 172. In some examples, the biasing element 172 is a pressure compensating spring it is housed in a housing 174. The pressure compensator valve 144 includes a drawer 170 which is designed to pass into the open position when the pump pressure at the pump outlet 128 exceeds a predetermined pressure, for example, when the pump pressure exceeds the predetermined pressure, the pump pressure at the pump outlet 128 acts on the slide 170 to override the biasing force from the biasing member 172 and to open the pressure compensator valve 144 providing communication to open between the pump outlet 128 and the pressure compensation line 142.
    The pressure compensation line 142 is connected to a pressure compensation actuator 136 which is adapted to receive hydraulic pressure from the pump outlet 128 when the pressure compensator valve 144 is opened. When hydraulic pressure is applied to the pressure compensation actuator 136, the pressure compensation actuator 136 changes from a rest position to an activated position causing the cylinder head 130 to move in a defusing direction from the position of maximum displacement D o towards position of minimum or zero displacement D f . The defusing of the cylinder head 130 by the pressure compensation actuator 136 means that the actual stroke length SL A of the pistons 112 is less than an authorized admissible stroke length SL MAX authorized by the swash plate 116 when the cylinder head 130 is located in the maximum displacement position D e .
    The slide 170 of the pressure compensator valve 144 may include a body 186 and a bearing 184. The diameter of the body 186 is less than the diameter of the bearing 184. When the pump pressure at the pump outlet 128 drops in below the predetermined pressure, the pressure compensator valve 144 switches to the closed position. When this occurs, the fluid from the pressure compensation line 142 flows into a tank (for example, the fluid reservoir 18 illustrated in Figure 1) by flowing through the pressure compensator valve 144. By example, when the slide 170 moves to the closed position, a space is formed between the bearing 184 and a surface of the pressure compensator valve 144 allowing the fluid to flow into a tank. When the fluid in the pressure compensation line 142 flows into a tank, the pressure compensation actuator 136 returns to the rest position and the cylinder head 130 returns to the maximum displacement position D c . because the cylinder head 130 is biased by a spring towards the maximum displacement position D o by the cylinder head return spring 162.
    As shown in Figure 4, the hydraulic pump system 12 further includes a speed compensator 148 with a speed compensation valve 152, a speed detection port 154 and a pump fixed displacement 158. The fixed displacement pump 158 is coupled to the drive shaft 106 and rotates together with the drive shaft 106 to provide a fixed displacement by rotation of the drive shaft 106. The pump fixed displacement 158 sucks hydraulic fluid from a pump inlet 126 of the variable displacement pump 102 and pumps the hydraulic fluid through port 154.
    The speed compensation valve 152 is biased by a biasing element 180 to a closed position where fluid communication between the pump outlet 128 and a speed compensation line 150 is blocked. In some examples, the element bias valve 180 is a speed compensator spring which is biased against one end of a slide 176 inside the speed compensation valve 152. In the closed position, fluid circulates through the orifice 154 in a pipe return 160 connected to one or more hydraulic fluid reservoirs or tanks such as the fluid reservoir 18 illustrated in FIG. 1. In certain examples, the flow of fluid in the return line 160 can be used to cool the various components of the hydraulic pump system 12 and others
    It systems.
    In low speed operation, the motor 146 can provide adequate torque to drive the rotary group 104 of the variable displacement pump 102 to the maximum displacement position D o . Consequently, the variable displacement pump 102 can operate at a maximum displacement by rotation of the rotary unit 104. In low speed operation, the speed compensator 148 is not activated and the displacement limitation actuator 156 is in rest position. However, the pressure compensator 134 can operate in low speed operation to move the cylinder head 130 in the defusing direction from the maximum displacement position D o to the minimum or zero displacement position D f to adjust the displacement of the pump to variable displacement 102 as a function of the pump pressure at the pump outlet 128.
    In some examples, the pressure compensator valve 144 is a distributor which controls the amount of pressure applied to the pressure compensation actuator 136 proportional to the pump pressure at the pump outlet 128. Thus, in operation at low speed, the position of the cylinder head 130 can be adjusted by the pressure compensation actuator 136 if necessary between the maximum displacement position D o and the minimum displacement position D ( .
    In high speed operation, the motor 146 cannot provide adequate torque to drive the drive shaft 106 and the rotary group 104 to the maximum displacement position D o . Thus, a new position of maximum displacement D N is defined by the speed compensator 148 to reduce the maximum admissible stroke length SL MAX of the pistons 112, so that the variable displacement pump 102 does not operate at maximum displacement by rotation of rotary group 104.
    Figure 5 is a schematic diagram of the hydraulic pump system 12 in operation, at high speed. As shown in FIG. 5, the speed compensation valve 152 is designed to pass into an open position when a differential pressure through the port 154 exceeds a predetermined value. The differential pressure across port 154 may increase due to increased flow from the fixed displacement pump 158. When the differential pressure across port 154 exceeds the predetermined value, pressure on an upstream side of the orifice 154 acts on an opposite end of the slide 176 by prevailing over the biasing force coming from the biasing element 180.
    The speed compensation valve 152 is a distributor provided with a variable opening which depends on the magnitude of the differential pressure through the orifice 154. A high differential pressure through the orifice 154 produces a more open large speed compensation valve 152 that a differential pressure [0059] [0060] [0061I [0062 low. The differential pressure across port 154 is proportional to the speed of the electric motor 146 as the increase in the speed of the electric motor increases the flow from the fixed displacement pump 158. Consequently, the extent to which the pressure compensation valve speed 152 is open is proportional to the speed of electric motor 146.
    The drawer 176 may include an annular groove 178 positioned between one or more guides 182 to connect a fluid from the pump outlet 128 to the speed compensation line 150 when the pressure on the upstream side of the orifice 154 causes the drawer 176 to pass in the open position.
    Hydraulic pressure from pump outlet 128 is applied to a displacement limiting actuator 156 via the speed compensation line 150 when the speed compensation valve 152 is in the open position. The displacement limiting actuator 156 can be biased towards a rest position by a biasing element 164 housed in a housing 166. When a fluid pressure circulates through the speed compensation line 150, the pressure prevails on the biasing force of the biasing element 164 to move the displacement limiting actuator 156 in a direction towards the minimum or zero displacement position D s . When activated, the displacement limiting actuator 156 moves the maximum displacement position D o of the cylinder head 130 in a direction towards the minimum or zero displacement position D b so that a new maximum displacement position D N is established and the maximum admissible stroke length SL MAX authorized by the swash plate 116 is reduced.
    The new maximum displacement position D N is positioned somewhere between the maximum displacement position D o and the minimum or zero displacement position Dp The new maximum displacement position D N reduces the torque required to drive the variable displacement pump 102 to the maximum admissible stroke length SL MAX . This can help prevent the motor 146 from stalling due to a high torque demand when the variable displacement pump 102 operates at the maximum allowable stroke length SL MAX .
    When the differential pressure across port 154 drops below the predetermined value due to a decrease in the flow rate from the fixed displacement pump 158 (resulting for example from a decrease in engine speed), the displacement limiting actuator 156 returns to the rest position, and a fluid coming from the speed compensation line 150 flows to the tank (for example, the fluid reservoir 18 illustrated in FIG. 1) by circulating at through an annular groove 168 of the slide 176 in the speed compensation valve 152. In addition, the swash plate 116 returns to the maximum displacement position D o due to the biasing force from the cylinder head return spring 162 .
    In high speed operation, the pressure compensator 134 can continue to operate to adjust the displacement of the variable displacement pump 102 as a function of the pump pressure at the pump outlet 128. However, in high speed operation , the pressure compensation actuator 136 disarms the cylinder head 130 between the new maximum displacement position D N and the minimum or zero displacement position Dp [0064] Figures 6 to 11 show the displacement limitation actuator 156 and l pressure compensation actuator 136 at different speeds and stroke lengths of the variable displacement pump 102. FIG. 6 represents the displacement limitation actuator 156 and the pressure compensation actuator 136 in operation at low speed and at a maximum admissible stroke length SL MAX . As shown in FIG. 6, the displacement limitation actuator 156 and the pressure compensation actuator 136 are both in their respective rest positions, and the variable displacement pump 102 operates at maximum displacement by rotation of the group rotary 104.
    FIG. 7 represents the displacement limitation actuator 156 and the pressure compensation actuator 136 in operation at low speed and at an intermediate stroke length. As shown in FIG. 7, the displacement limitation actuator 156 remains at rest, however, the pressure compensation actuator 136 is activated (for example, because the pump pressure at the pump outlet 128 exceeds the predetermined pressure), and the pressure compensation actuator 136 moves in the direction towards the minimum or zero displacement position D s to defuse the cylinder head 130 and reduce the actual stroke length SL A of the pistons 112, La FIG. 8 represents the displacement limitation actuator 156 and the pressure compensation actuator 136 in operation at medium speed and at a maximum stroke length SL MAX . As shown in FIG. 8, the displacement limiting actuator 156 creates a new maximum displacement position In a direction towards the minimum or zero displacement position Dp The pressure compensation actuator 136 is in a rest position because the new maximum displacement position D N decreases the maximum admissible stroke length SL MA x of the pistons 112.
    Figure 9 shows the displacement limitation actuator 156 and the pressure compensation actuator 136 in operation at medium speed and at an intermediate stroke length. As shown in FIG. 9, the displacement limitation actuator 156 and the pressure compensation actuator 136 are both activated to reduce the actual stroke length SL A of the pistons 112.
    10 shows the displacement limitation actuator 156 and the pressure compensation actuator 136 in operation at high speed and at a maximum stroke length SL MAX . As shown in FIG. 10, the displacement limiting actuator 156 is activated so as to establish a new maximum displacement position D N + i in a direction closer to the minimum or zero displacement position D <. The pressure compensation actuator 136 is in a quiescent state. However, due to the new maximum displacement position D N + 1 , the maximum permissible stroke length SL MAX of the pistons 112 has further decreased compared to the maximum permissible stroke length SL MAX during the medium speed operation illustrated in FIG. figure 8.
    FIG. 11 represents the displacement limitation actuator 156 and the pressure compensation actuator 136 in operation at high speed and at an intermediate stroke length. As shown in Figure 11, the displacement limiting actuator 156 and the pressure compensation actuator 136 are both activated to reduce the actual stroke length SL A of the pistons 112. Due to the new maximum displacement position D N + i in a direction closer to the minimum or zero displacement position D <, the actual stroke length SL A of the pistons 112 has further decreased compared to the stroke length during the medium speed operation illustrated in the figure 9.
    Figure 12 is a graph illustrating a maximum allowable stroke with respect to an electric motor speed. As shown in FIG. 12, a predetermined speed threshold S Po is defined as the speed limit at which the electric motor 146 can operate safely at the maximum displacement position D o and at the maximum authorized stroke SL MAX . When the engine speed 146 reaches the predetermined speed threshold SP 0 , the displacement limitation actuator 156 is activated to move the maximum displacement position D o of the cylinder head 130 in a direction towards the minimum or zero displacement position Dp This establishes a new maximum displacement position D N for the cylinder head 130 which is closer to the minimum or zero displacement position D · ,, As described above, the speed compensation valve 152 of the speed compensator 148 is a distributor, the size of the opening of which depends on the magnitude of the differential pressure through the orifice 154. Thus, when the engine speed 146 increases, the position of maximum displacement D N is moved closer towards the minimum or zero displacement position Dj so that the maximum admissible stroke length SL MAX decreases when the engine speed 146 increases beyond the predetermined speed threshold S P 0 . In this way, the speed compensator 148 can prevent the motor 146 from being submerged by controlling the maximum admissible stroke length SL MAX , and therefore the torque required to drive the pistons 112 of the rotary unit 104 to the maximum admissible stroke length SL MAX .
    Furthermore, the pressure compensator 134 operates at speeds both lower and higher than the predetermined speed threshold SP 0 to adjust the stroke length of the pistons 112 as a function of the pump pressure at the pump outlet 128. Thus, the actual stroke length SL A of the pistons 112 of the variable displacement pump 102 is defined in the area of the curve illustrated in FIG. 12.
    When the hydraulic pump system 12 is installed in an aircraft, the pressure compensator 134 can frequently defuse the cylinder head 130 during the flight of the aircraft because consumers (for example, downstream components 16) of the aircraft 13 are active during the flight, and given the need for the hydraulic pump system 12 to maintain a constant pressure for consumers (for example, downstream components 16). On the contrary, the speed compensator 148 will not frequently adjust the maximum displacement position D o of the cylinder head 130 during the flight of the aircraft 13 because the speed compensator 148 responds to changes in engine speed 146 which are determined , at least in part, by the engine speed of the aircraft. Although the aircraft engine speed is generally high at take-off and low at landing, the aircraft engine speed generally does not change frequently at cruising altitude. Thus, the speed compensator 148 will generally not be as active as the pressure compensator 134 during the flight of the aircraft.
    Returning to Figures 4 and 5, the hydraulic pump system 12 further includes a negative cylinder head moment actuator 188. The negative cylinder head moment actuator 188 applies additional force to the cylinder head 130 in one direction towards the minimum or zero displacement position D [. The additional force compensates for a positive momentary force exerted on the cylinder head 130 which results from an increase in the speed of the drive shaft 106 driven by the electric motor 146.
    In certain examples, the negative torque momentary actuator 188 is biased towards a rest position by a biasing element 196 housed in a housing 192. When the differential pressure across the orifice 154 exceeds a predetermined value ( for example due to an increased flow coming from the fixed displacement pump 158 which results from an increased speed of the drive shaft 106), a pressure circulates through the speed compensation line 150. The pressure of fluid flowing through the speed compensation line 150 wins over the biasing force of the biasing element 196 to move the negative cylinder head momentum actuator 188 in a direction towards the minimum or zero displacement position Dj. When activated, the negative yoke moment actuator 188 pushes a negative yoke moment spring 190 to apply a negative moment force on the yoke 130 of the variable displacement pump 102. The negative yoke moment 190 is positioned between the cylinder head 130 and a surface 198 of the negative cylinder head actuator 188.
    As described above, the fluid in the speed compensation line 150 flows through the annular groove 168 of the slide 176 when the speed compensation valve 152 passes into the closed position. When the differential pressure across port 154 drops below the predetermined value due to a decrease in the flow rate from the fixed displacement pump 158 (resulting for example from a decrease in engine speed), the negative cylinder head actuator 188 returns to the rest position due to the biasing force from the biasing member 196.
    Figures 13 to 18 show the negative cylinder head momentum actuator 188 at different speeds and stroke lengths of the variable displacement pump 102. Figure 13 shows the negative cylinder head momentum actuator 188 in low operation speed and maximum permissible stroke length SL MAX . As shown in FIG. 13, the negative cylinder head moment actuator 188 is in a rest position when the variable displacement pump 102 operates at low speed and at the maximum admissible stroke length SL MAX . The negative cylinder head spring 190 is also in a state of rest.
    FIG. 14 represents the negative cylinder head moment actuator 188 in operation at low speed and at an intermediate stroke length. As shown in FIG. 14, the negative cylinder head actuator 188 remains in the rest position when the variable displacement pump 102 operates at low speed and at an intermediate stroke length. However, during this mode of operation, the negative cylinder head moment spring 190 is decompressed in one direction to the minimum or zero displacement position D [because the cylinder head 130 has been moved in this direction. The cylinder head 130 is moved to the minimum or zero displacement position D, by the pressure compensation actuator 136 when the pump pressure at the pump outlet 128 exceeds the predetermined pressure.
    FIG. 15 represents the negative cylinder head moment actuator 188 in operation at medium speed and at the maximum admissible stroke length SL MAX . As shown in Figure 15, the negative cylinder head momentum actuator 188 is compressed in one direction to the minimum or zero displacement position Di. During this mode of operation, the negative yoke moment spring 190 applies a negative moment force to the yoke 130 which compensates for the positive moment force applied to the yoke 130 due to the increased speed of the shaft d drive 106 and the electric motor 146.
    FIG. 16 represents the negative cylinder head moment actuator 188 in operation at medium speed and at an intermediate stroke length. As shown in Figure 16, the cylinder head 130 is moved in a direction towards the minimum or zero displacement position D, relative to the position of the cylinder head 130 shown previously in Figure 15. The cylinder head 130 is moved to the position of minimum or zero displacement Dt by the pressure compensation actuator 136 when the pump pressure at the pump outlet 128 exceeds the predetermined pressure. During this operating mode, the negative yoke moment actuator 188 and the surface 198 remain in the same position as in FIG. 15; however, the negative cylinder head moment spring 190 is decompressed in one direction to the minimum or zero displacement position Djdu that the cylinder head 130 has been moved in this direction.
    FIG. 17 represents the negative cylinder head moment actuator 188 in operation at high speed and at the maximum stroke length SLmax- As shown in FIG. 17, the negative cylinder head moment actuator 188 is compressed in a direction towards the minimum or zero displacement position Dj, and a greater negative moment force is applied by the negative cylinder head spring 190.
    Figure 18 shows the negative cylinder head momentum actuator 188 in operation at high speed and at an intermediate stroke length. As shown in FIG. 18, the cylinder head 130 moves in a direction towards the position of minimum or zero displacement Ο, with respect to the position of the cylinder head 130 in FIG. 17. As described above, the cylinder head 130 can be moved to the minimum or zero displacement position D> by the pressure compensation actuator 136 when the pump pressure at the pump outlet 128 exceeds the predetermined pressure. During this operating mode, the negative cylinder head momentary actuator 188 and the surface 198 remain in the same position as in FIG. 17; however, the negative cylinder head moment spring 190 is decompressed in one direction to the minimum or zero displacement position D [because the cylinder head 130 has been moved in this direction.
    Figure 19 is a graph illustrating a cylinder head moment with respect to an electric motor speed. As shown in FIG. 19, the curve M P illustrates the positive momentary force applied to the cylinder head 130 when the speed (rpm) of the drive shaft 106 increases. The curve illustrates the negative moment force applied by the negative yoke moment actuator 188 when the speed of the drive shaft 106 increases. As shown, the normal increase in the cylinder head moment as the speed increases is compensated by the moment imposed by the negative cylinder head actuator 188. Consequently, the line M o illustrates the cylinder head moment over which the displacement limitation actuator 156 when the speed of the drive shaft 106 increases.
    [0085]
    0086 [[0087] [0088] [0089]
    FIG. 20 is a functional diagram illustrating a method 600 for controlling the displacement of a variable displacement pump. Referring now to Figure 20, the method 600 includes a step 602 of establishing a maximum displacement position for a pivoting cylinder head supporting a swash plate in a variable displacement pump driven by an electric motor.
    The method 600 includes a step 604 consisting in determining if the electric motor speed exceeds a predetermined speed threshold. When the predetermined speed threshold is exceeded, the method 600 includes a step 606 consisting in actuating a displacement limiting actuator to pivot the cylinder head in a new position of maximum displacement in a direction towards a position of minimum or zero displacement so that '' a maximum admissible stroke length is reduced.
    The method 600 includes a step 608 of determining whether the pump pressure at a pump outlet of the variable displacement pump exceeds a predetermined pressure. When the predetermined pressure is exceeded, the method 600 can further include a step 610 consisting in defusing the cylinder head in the direction towards the minimum or zero displacement position so that the displacement pump does not operate at the maximum admissible stroke length.
    In some examples, method 600 may further include a step 612 of applying a negative moment force to the cylinder head to compensate for a positive moment force on the cylinder head resulting from an increase in the speed at which the cylinder head rotates the drive shaft axis. In some examples, step 612 can take place at the same time as step 606 so that a negative moment force is applied while the cylinder limiting actuator rotates the cylinder head to the new maximum cylinder position. .
    In some examples, step 604 includes determining if a pressure differential across an orifice exceeds a predetermined value to determine if the electric motor speed exceeds the predetermined speed threshold. In some examples, step 606 includes opening a speed compensation valve to apply hydraulic pressure from a pump outlet of the variable displacement pump to the displacement limitation actuator. In some examples, the opening of the speed compensation valve in step 606 may vary depending on the magnitude of the differential pressure through the orifice which is proportional to the speed of the electric motor.
    The various embodiments described above are provided by way of illustration only and should not be interpreted as limiting the claims appended hereto. Those skilled in the art will readily identify various modifications and changes that can be made without following the examples of embodiments and applications illustrated and described herein, and without departing from the spirit and
    权利要求:
    Claims (1)
    [1" id="c-fr-0001]
    true scope of the following claims.
    Claim 1] [Claim 2] [Claim 3] [Claim 4]
    Claim 5] [Claim 6]
    Claim 7]
    claims
    Pump system comprising:
    a variable displacement pump including an inlet and an outlet, the variable displacement pump also including a drive shaft; an electric motor for driving a rotation of the drive shaft;
    a pressure compensator for adjusting a pump displacement of the variable displacement pump based on a pump pressure at the pump outlet; and a speed compensator for adjusting a maximum amplitude of the pump displacement inversely proportional to a speed of rotation of the drive shaft once the speed of rotation of the drive shaft has exceeded a predetermined speed threshold . The pump system of claim 1, wherein the pressure compensator includes a pressure compensating actuator adapted to receive hydraulic pressure from the pump outlet when a pressure compensator valve is opened.
    The pump system of claim 2, wherein the pressure compensation actuator defuses a cylinder head supporting a swash plate in the variable displacement pump when the pump pressure exceeds a predetermined pressure.
    The pump system of claim 1, wherein the speed compensator includes a displacement limiting actuator adapted to receive hydraulic pressure from the pump outlet when a speed compensation valve is opened. Pump system according to claim 4, in which the speed compensation valve is a distributor which has a variable opening proportional to the speed of the electric motor.
    The pump system of claim 5, wherein a speed compensator includes a fixed displacement pump coupled to the drive shaft and adapted to suck hydraulic fluid from the inlet of the variable displacement pump and to pump hydraulic fluid through an orifice.
    Pump system according to claim 6, wherein a pressure differential across the orifice is proportional to the speed of the electric motor and, when the pressure differential across the orifice exceeds a predetermined value, pressure on an upstream side from [Claim 8] [Claim 9] [Claim 10]
    Claim 11] [Claim 12] [Claim 13 the orifice opens the speed compensation valve.
    The pump system of claim 6, wherein the hydraulic fluid flows through the port in a return line for cooling components of the pump system.
    The pump system of claim 1, wherein the variable displacement pump includes a rotary group coupled to the drive shaft so that the rotary group rotates together with the drive shaft about a shaft axis drive, the rotary group including a piston block defining a plurality of cylinder bores positioned around the axis of the drive shaft, the rotary group being provided with pistons which perform a reciprocating movement in the cylinder bores to pump hydraulic fluid through the variable displacement pump from the inlet to the outlet, the pistons being provided with piston shoes.
    The pump system of claim 9, wherein the variable displacement pump further includes a cylinder head supporting a swash plate, the piston shoes of the rotary group pistons being configured to slide along the swash plate while the rotary group rotates around of the drive shaft axis, the cylinder head being able to move in a votive manner with respect to the drive shaft axis to adjust a swash plate angle of the swash plate with respect to the axis of drive shaft.
    The pump system of claim 10, wherein the speed compensator includes a displacement limiting actuator adapted to move a maximum displacement position of the cylinder head to reduce a maximum allowable stroke length allowed by the swash plate.
    The pump system of claim 11, wherein the pressure compensator includes a pressure compensating actuator adapted to defuse the cylinder head when a pump pressure at the outlet of the variable displacement pump exceeds a predetermined pressure so that the length stroke of the rotary unit pistons is less than the maximum admissible stroke length authorized by the swash plate.
    The pump system of claim 10, further comprising a negative cylinder head actuator which applies a negative moment force to compensate for a positive moment force on the cylinder head.
    [Claim 14]
    Claim 15] [Claim 16 [Claim 17]
    Claim 18] [Claim 19]
    The pump system of claim 1, wherein the electric motor is adapted to receive variable frequency electrical energy from an energy source on an aircraft.
    Method for controlling a pump displacement of a variable displacement pump, the variable displacement pump being provided with a cylinder head supporting a swash plate, the cylinder head being driven by an electric motor to rotate about an axis of drive shaft and the cylinder head pivotally movable relative to the drive shaft axis, the method comprising the steps of: establishing a maximum displacement position for the cylinder head;
    determining if the electric motor speed exceeds a predetermined speed;
    when the electric motor speed exceeds the predetermined speed, actuate a displacement limiting actuator to pivot the cylinder head towards a minimum or zero displacement position to establish a new maximum displacement position which reduces a maximum admissible stroke length authorized by the swash plate; determining if the pump pressure at a pump outlet of the variable displacement pump exceeds a predetermined pressure; and when the pump pressure exceeds the predetermined pressure, defuse the cylinder head towards the minimum or zero displacement position so that the stroke length authorized by the swash plate is less than the maximum admissible stroke length.
    The method of claim 15, wherein the step of determining whether the electric motor speed exceeds the predetermined speed is performed by determining whether a differential pressure through an orifice exceeds a predetermined value.
    The method of claim 16, wherein the actuation of the displacement limiting actuator is effected by opening a speed compensation valve for applying hydraulic pressure from the pump output of the variable displacement pump to the actuation of displacement limitation.
    The method of claim 17, wherein the size of the opening of the speed compensation valve varies as a function of the magnitude of the differential pressure through the orifice which is proportional to the speed of the electric motor.
    The method of claim 15, further comprising applying a negative moment force to the cylinder head to compensate for a positive moment force [Claim 20] on the cylinder head resulting from an increase in the speed at which the cylinder head rotates of the drive shaft axis. A pump system comprising: a variable displacement pump including: a pump housing defining a pump inlet and a pump outlet;
    a drive shaft which extends through the pump housing, the drive shaft being rotatable about a drive shaft axis;
    a rotary group inside the pump housing, the rotary group being coupled to the drive shaft so that the rotary group rotates together with the drive shaft about the drive shaft axis , the rotary group including a piston block defining a plurality of cylinder bores positioned around the axis of the drive shaft, the rotary group also including pistons provided with piston bodies which reciprocate - comes in the cylinder bores to pump hydraulic fluid through the in-line pump from the pump inlet to the pump outlet, the pistons also including piston shoes;
    a cylinder head inside the pump housing, the cylinder head supporting a swash plate, the piston shoes of the rotary group pistons being configured to slide along the swash plate while the rotary group rotates around the shaft shaft drive, the cylinder head being movable pivotally relative to the drive shaft axis to adjust a swash plate angle of the swash plate relative to the drive shaft axis, the angle of the plate oscillating establishing a displacement of the pump of the variable displacement pump by rotation of the rotary group by establishing a length of stroke of the pistons in the cylinder bores, the cylinder head being able to move in a rotary manner with respect to the shaft axis drive between a maximum displacement position and a minimum or zero displacement position, the pistons having a maximum stroke length when the cylinder head is in the position maximum displacement and a minimum stroke length or no stroke length when the cylinder head is in the minimum or zero displacement position, the cylinder head being biased by a spring towards the maximum displacement position;
    a pressure compensator for adjusting the pump displacement based on a pump pressure at the pump outlet, the pressure compensator including a pressure compensation actuator for moving the cylinder head in a defusing direction from the displacement position towards the minimum or zero displacement position, the pressure compensator also including a pressure compensation line for applying hydraulic pressure to the pressure compensation actuator to move the cylinder head in the defusing direction, the pressure compensator also including a pressure compensator valve for opening and closing a fluid communication between the pressure compensation line and the pump outlet, the pressure compensator valve being movable between a closed position in which a fluid communication is blocked between the pump outlet and compensation line pressure n and an open position in which a fluid communication is open between the pump outlet and the pressure compensation line, the pressure compensator valve being biased towards the closed position by a biasing force, the pressure compensator valve pressure being configured to open when the pump pressure at the pump outlet exceeds the predetermined pressure;
    an electric motor for driving a rotation of the drive shaft;
    a speed compensator including:
    an orifice;
    a fixed displacement pump coupled to the drive shaft so that the drive shaft rotates the fixed displacement pump, the fixed displacement pump being configured to draw hydraulic fluid from one side d input of the variable displacement pump and for pumping hydraulic fluid through the orifice;
    a speed compensation line for applying hydraulic pressure to the displacement limiting actuator to move the maximum cylinder displacement position to the minimum or zero displacement position, so that a new maximum displacement position is established and that a maximum admissible stroke length authorized by the swash plate is reduced;
    a speed compensation valve for opening and closing a fluid communication between the pump outlet and the speed compensation line, the speed compensation valve being biased towards a closed position where a fluid communication between the pump outlet and the line speed compensation valve is blocked, the speed compensation valve being configured to open fluid communication between the pump outlet and the speed compensation line when a differential pressure across the port exceeds a predetermined value.
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    BE520187A|
    同族专利:
    公开号 | 公开日
    US20190145390A1|2019-05-16|
    引用文献:
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    法律状态:
    2019-10-22| PLFP| Fee payment|Year of fee payment: 2 |
    2020-10-21| PLFP| Fee payment|Year of fee payment: 3 |
    2021-07-30| PLSC| Publication of the preliminary search report|Effective date: 20210730 |
    2021-10-20| PLFP| Fee payment|Year of fee payment: 4 |
    优先权:
    申请号 | 申请日 | 专利标题
    US201762585855P| true| 2017-11-14|2017-11-14|
    US62/585855|2017-11-14|
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