ADJUSTABLE AIR GAP AXIAL FLUX MOTOR

An axial flux motor may include a stator, a rotor including a rotor frame and having a rotational axis, an axial air gap separating the stator and the rotor, the axial air gap corresponding to an axial position of the rotor frame, and at least one hydraulic actuator including a variable volume hydraulic fluid chamber defined within the rotor frame, wherein a change in the volume of the variable volume hydraulic fluid chamber causes a change in the axial position of the rotor frame and a corresponding change in the axial air gap separating the stator and the rotor.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application claims priority to Chinese Patent Application No. 202211498240.X, filed Nov. 28, 2022, the contents of which are incorporated by reference herein in their entirety.

INTRODUCTION

The subject disclosure relates to permanent magnet (PM) axial flux (AF) motors. PM motors may produce back-EMF within the stator windings proportional to the motor speed and air gap field strength. Without some type of field weakening above base speed operation, a PM motor may be speed and torque limited.

SUMMARY

In one exemplary embodiment, an axial flux motor may include a stator, a rotor including a rotor frame and having a rotational axis, an axial air gap separating the stator and the rotor, the axial air gap corresponding to an axial position of the rotor frame, and at least one hydraulic actuator including a variable volume hydraulic fluid chamber defined within the rotor frame, wherein a change in the volume of the variable volume hydraulic fluid chamber causes a change in the axial position of the rotor frame and a corresponding change in the axial air gap separating the stator and the rotor.

In addition to one or more of the features described herein, the at least one hydraulic actuator may be a linear hydraulic actuator.

In addition to one or more of the features described herein, the at least one hydraulic actuator may be a rotary hydraulic actuator.

In addition to one or more of the features described herein, the axial flux motor may further include a motor output shaft having a passage communicating a hydraulic fluid to and from the variable volume hydraulic fluid chamber.

In addition to one or more of the features described herein, the at least one hydraulic actuator may include a plurality of variable volume hydraulic fluid chambers.

In addition to one or more of the features described herein, the at least one hydraulic actuator may include a plurality of hydraulic actuators angularly distributed around the rotor.

In addition to one or more of the features described herein, the axial flux motor may further include a motor output shaft having a passage communicating hydraulic fluid to the variable volume hydraulic fluid chamber, wherein the variable volume hydraulic fluid chamber may include a cavity in the rotor frame enclosing a vein plate, the vein plate extending radially outward from a cylindrical base surrounding the motor output shaft, the cylindrical base affixed to the motor output shaft such that the cylindrical base is rotationally and axially fixed relative to the motor output shaft, the rotor frame affixed to the cylindrical base such that the rotor frame is axially translatable.

In addition to one or more of the features described herein, the axial flux motor may further include a motor output shaft including a passage communicating hydraulic fluid to the variable volume hydraulic fluid chamber, wherein the variable volume hydraulic fluid chamber may include a cavity in the rotor frame enclosing a vein plate, the vein plate extending radially outward from the motor output shaft, the vein plate affixed to the motor output shaft such that the vein plate is rotationally fixed relative to the motor output shaft and axially translatable relative to the motor output shaft, the rotor frame affixed to the motor output shaft such that the rotor frame is rotationally and axially translatable relative to the motor output shaft.

In addition to one or more of the features described herein, the rotor frame may be affixed to the motor output shaft by a lead screw coupling.

In addition to one or more of the features described herein, the lead screw coupling may include a ball screw.

In addition to one or more of the features described herein, the axial flux motor may include a disc spring compressible during the change in the axial position of the rotor frame.

In addition to one or more of the features described herein, the axial flux motor may further include a binary fluid control valve having a first state providing the hydraulic fluid to the hydraulic fluid chamber and a second state exhausting the hydraulic fluid from the hydraulic fluid chamber.

In addition to one or more of the features described herein, the rotor may include surface mounted permanent magnets.

In another exemplary embodiment, an axial flux motor may include a stator, a rotor including a rotor frame and having a rotational axis, an axial air gap separating the stator and the rotor, the axial air gap corresponding to an axial position of the rotor frame, a motor output shaft, at least one hydraulic actuator including a variable volume hydraulic fluid chamber defined within the rotor frame, the variable volume hydraulic fluid chamber may include a cavity in the rotor frame enclosing a vein plate, the vein plate extending radially outward from a cylindrical base surrounding the motor output shaft, the cylindrical base affixed to the motor output shaft such that the cylindrical base is rotationally and axially immovable relative to the motor output shaft, the rotor frame affixed to the cylindrical base such that the rotor frame is axially translatable, wherein a change in the volume of the variable volume hydraulic fluid chamber causes a change in the axial position of the rotor frame and a corresponding change in the axial air gap separating the stator and the rotor, the motor output shaft communicating a hydraulic fluid to and from the variable volume hydraulic fluid chamber, and a binary fluid control valve having a first state providing the hydraulic fluid to the hydraulic fluid chamber and a second state exhausting the hydraulic fluid from the hydraulic fluid chamber.

In addition to one or more of the features described herein, the axial flux motor may further include a spring biasing the rotor away from the stator when compressed.

In addition to one or more of the features described herein, the axial flux motor may further include a spring biasing the rotor toward from the stator when compressed.

In yet another exemplary embodiment, an axial flux motor may include a rotational axis of the motor, a stator having a pair of axial sides, a rotor including a respective rotor structure on each axial side of the stator, each rotor structure including a rotor frame, a respective axial air gap separating the stator and each respective rotor structure, each axial air gap corresponding to an axial position of the corresponding rotor frame, for each rotor structure, at least one respective hydraulic actuator may include a corresponding variable volume hydraulic fluid chamber defined within the respective rotor frame, wherein a change in the volume of the corresponding variable volume hydraulic fluid chamber causes a change in the axial position of the respective rotor frame and a corresponding change in the corresponding axial air gap separating the stator and the respective rotor structure, and a binary fluid control valve having a first state providing a hydraulic fluid to each variable volume hydraulic fluid chamber and a second state exhausting the hydraulic fluid from each variable volume hydraulic fluid chamber.

In addition to one or more of the features described herein, for each rotor structure the at least one respective hydraulic actuator may be a linear hydraulic actuator.

In addition to one or more of the features described herein, for each rotor structure the at least one respective hydraulic actuator may be a rotary hydraulic actuator.

In addition to one or more of the features described herein, for each rotor structure the corresponding variable volume hydraulic fluid chamber may include a respective cavity in the respective rotor frame enclosing a respective vein plate, the respective vein plate extending radially outward from a respective cylindrical base surrounding a motor output shaft, the respective cylindrical base affixed to the motor output shaft such that the respective cylindrical base is rotationally and axially fixed relative to the motor output shaft, the respective rotor frame affixed to the respective cylindrical base such that the respective rotor frame is axially translatable.

DETAILED DESCRIPTION

FIG.1schematically illustrates an embodiment of an electric propulsion system101in a vehicle100. Vehicle and vehicular are understood to refer to any means of transportation including non-limiting examples of motorcycles, cars, trucks, buses, excavation, earth moving, construction and farming equipment, railed vehicles like trains and trams, and watercraft like ships and boats. The electric propulsion system101may include various control components, electrical systems and electro-mechanical systems including, for example, a rechargeable energy storage system (RESS)104and an electric drive unit (EDU)102. The electric propulsion system101may be employed on a powertrain system to generate propulsion torque as a replacement for, or in conjunction with, an internal combustion engine in various electric vehicle (EV) applications and hybrid electric vehicle (HEV) applications, respectively.

The EDU102may be of varying complexity, componentry and integration. An exemplary highly integrated EDU102may include, for example, a rotary electric machine such as an alternating current (AC) motor (motor)120and a traction power inverter module (TPIM)106including a motor controller105and a power inverter110. The motor120may include a stator120S and a rotor120R coupled to a motor output shaft125and position sensor182, for example a variable reluctance resolver or an encoder. The position sensor182may signally connect directly to the motor controller105and is employed to monitor angular position of the rotor (θe) of the motor120. The angular position of the rotor (θe) of the motor120is employed by the motor controller105to control operation of the power inverter110that controls the motor120.

The motor output shaft125may transfer torque between the motor120and driveline components (not illustrated), some of which may be integrated within the EDU102, for example in a gearbox including reduction and differential gear sets and one or more axle outputs. The gearbox may simply include reduction gearing and a prop shaft output for coupling to a differential gear set. One or more axles may couple to the gear box directly or through final drive or differential gear sets if separate therefrom. Axle(s) may couple to a vehicle wheel(s) for transferring tractive force between a wheel and pavement. One having ordinary skill in the art will recognize alternative arrangements for driveline components. Propulsion torque requests or commands136(Tcmd) may be provided by a vehicle controller103to the motor controller105.

Any controller may include one or more control modules. As used herein, control module, module, control, controller, control unit, electronic control unit, processor and similar terms mean any one or various combinations of one or more of Application Specific Integrated Circuit(s) (ASIC), electronic circuit(s), central processing unit(s) (preferably microprocessor(s)) and associated memory and storage (read only memory (ROM), random access memory (RAM), electrically programmable read only memory (EPROM), hard drive, etc.) or microcontrollers executing one or more software or firmware programs or routines, combinational logic circuit(s), input/output circuitry and devices (I/O) and appropriate signal conditioning and buffer circuitry, high speed clock, analog to digital (A/D) and digital to analog (D/A) circuitry and other components to provide the described functionality. A control module may include a variety of communication interfaces including point-to-point or discrete lines and wired or wireless interfaces to networks including wide and local area networks, and in-plant and service-related networks including for over the air (OTA) software updates. Functions of a control module as set forth in this disclosure may be performed in a distributed control architecture among several networked control modules. Software, firmware, programs, instructions, routines, code, algorithms and similar terms mean any controller executable instruction sets including calibrations, data structures, and look-up tables. A control module may have a set of control routines executed to provide described functions. Routines are executed, such as by a central processing unit, and are operable to monitor inputs from sensing devices and other networked control modules and execute control and diagnostic routines to control operation of actuators. Routines may be executed at regular intervals during ongoing powertrain and vehicle operation. Alternatively, routines may be executed in response to occurrence of an event, software calls, or on demand via user interface inputs or requests.

The RESS104may, in one embodiment, include one or more electro-chemical battery packs112, for example high capacity, high voltage (HV) rechargeable lithium ion battery packs for providing power to the vehicle via a HV direct current (DC) bus108. The RESS104may also include a battery manager module114. The RESS104may include one or more battery packs112constructed from a plurality of battery pack modules allowing for flexibility in configurations and adaptation to application requirements. Battery packs may include a plurality of battery pack modules constructed from a plurality of cells allowing for flexibility in configurations and adaptation to application requirements. Battery pack modules may include a plurality of cells allowing for flexibility in configurations and adaptation to application requirements. For example, in vehicular uses, the RESS104may be modular to the extent that the number of battery pack modules may be varied to accommodate a desired energy density or range objective of a particular vehicle platform, intended use, or cost target. Battery packs and battery pack modules may be variously and selectively configured in accordance with desired propulsion architecture and charging functions. It is understood that the RESS104may be reconfigurable at any level of integration including battery pack, battery module and cell.

The motor120may be a multi-phase AC motor receiving multi-phase AC power over a multi-phase motor control power bus (AC bus)111which is coupled to the power inverter110. In one embodiment, the motor120is a three-phase motor and the power inverter110is a three-phase inverter. The power inverter110may include a plurality of solid-state switches in a solid-state switching section. The power inverter110couples to DC power over the HV DC bus108(DC input voltage (Vdc)) from the RESS104, for example at 400 or 800 volts. The motor controller105is coupled to the power inverter110for control thereof. The power inverter110electrically connects to stator phase windings of a three-phase stator winding of the motor120via the AC bus111, with electric current (Iabc) monitored on two or three phases thereof. The power inverter110may be configured with suitable control circuits including paired power transistors (e.g., IGBTs) for transforming high-voltage DC voltage on the HV DC bus108to high-voltage three-phase AC voltage (Vabc) on the AC bus111and transforming high-voltage three-phase AC voltage (Vabc) on the AC bus111to high-voltage DC voltage on the HV DC bus108. The power inverter110may employ any suitable pulse width modulation (PWM) control, for example sinusoidal pulse width modulation (SPWM) or space vector pulse width modulation (SVPWM), to generate switching vector signals (Sabc)109to convert stored DC electric power originating in the battery pack112of the RESS104to AC electric power to drive the motor120to generate torque. Similarly, the power inverter110may convert mechanical power transferred to the motor120to DC electric power to generate electric energy that is storable in the battery pack112of the RESS104, including as part of a regenerative braking control strategy. The power inverter110may be configured to receive the switching vector signals (Sabc)109from motor controller105and control inverter states to provide the motor drive and regeneration functionality. Switching vector signals (Sabc)109may also be referred to herein as conduction commands.

Control of the power inverter110may include high frequency switching of the solid-state switches in accordance with the PWM control. A number of design and application considerations and limitations determine inverter switching frequency and PWM control. Inverter controls for AC motor applications may include fixed switching frequencies, for example switching frequencies around 10-30 kHz and PWM controls that minimize switching losses of the IGBTs or other power switches of the power inverter110.

The motor120in the EDU102of the electric propulsion system101in one embodiment may be a permanent magnet axial flux (PMAF) motor. The disclosed improvement relates to mechanically based field weakening apparatus and control of such a motor.FIGS.2A and2Bschematically illustrates one embodiment of a PMAF motor220in pertinent parts andFIGS.3A and3Bschematically illustrates another embodiment of a PMAF motor220in pertinent parts. The motor220may include a stator220S, a rotor220R, and a motor output shaft225. A motor frame230may support the stator220S and bearings240. The bearings240in turn rotatably support the motor output shaft225. Each bearing240may include an outer ring240A affixed to the motor frame230, rolling elements (e.g., balls, rollers, etc.), and an inner ring240B affixed to the motor output shaft225, for example through mounting hardware such as an inner ring extension/concentric collar or a tapered adapter sleeve/locking washer/nut complement. The stator220S may be unitary including a common core or may be segmented and coreless. The rotor220R may include rotor structure on one or both axial sides of the stator220S. In one embodiment as illustrated, the rotor220R includes rotor structure on both axial sides of the stator220S. In another embodiment, the rotor220R includes rotor structure on just one axial side of the stator220S. Further description herein will refer to one rotor structure on one side of the stator220S for ease of description and understanding. The rotor structure may include a rotor frame251, rotor core253and permanent magnets255. The rotor core253may be affixed to the rotor frame251for rotation therewith. Permanent magnets255may be affixed to the rotor core253for rotation therewith and with the rotor frame251. The rotor frame251may be formed from a soft magnetic composite material for example and the rotor core253may be made from electric steel laminations for example. In the illustrated embodiments ofFIGS.2A/2B and3A/3B, the permanent magnets are surface mounted to the rotor core253and arranged between the rotor core253and the stator220S with air gaps260therebetween. In alternate embodiments, the permanent magnets255may be contained within the rotor core253with the air gaps260between the rotor core253and the stator220S.

The rotor220R may be movable along the rotational axis (A) of the rotor220R. Thus, the air gaps260A/260B may be varied. In the illustrated embodiments,FIGS.2A and3Ashow first relatively narrow air gaps260A whereasFIGS.2B and3Bshow second relatively wide air gaps260B. The relatively narrow air gaps260A may represent a minimum attainable air gap for the motor within its mechanical design constraints, for example 1 mm. The relatively wide air gaps260B may represent a maximum attainable air gap for the motor within its mechanical design constraints, for example 2 mm. In each of the embodiments 2A/2B and 3A/3B, axial displacement of the rotor220R along the rotational axis (A) may be controlled by hydraulic fluid within variable volume hydraulic fluid chambers as further described herein.

In the embodiment ofFIGS.2A and2B, the rotor220R may include at least one hydraulic fluid chamber270within the rotor frame251.FIGS.2A and2Billustrate an embodiment including two hydraulic fluid chambers270stacked axially.FIGS.2A and2Balso only illustrate one hydraulic actuator in section at one angular orientation of the rotor220R, it being understood that multiple such hydraulic actuators may be angularly distributed around the entire rotor220R for system balance requirements and overall hydraulic force multiplication proportional to the number of hydraulic actuators and hydraulic fluid chambers which may advantageously allow for lower system hydraulic pressures. Each hydraulic fluid chamber270may include a cavity272formed in the rotor frame251and a vein plate271extending radially outward from a cylindrical base273which surrounds the motor output shaft225. The cylindrical base273is affixed to the motor output shaft225such that it is rotationally and axially immovable relative to the motor output shaft225. Thus, the vein plate271of the hydraulic fluid chamber is rotationally and axially static with respect to the motor output shaft and rotates therewith. The rotor frame251is affixed to the cylindrical base273such that it is axially translatable. The cavity272in the rotor frame251encloses the vein plate271and together therewith defines the hydraulic fluid chamber270. Thus, the vein plate271and the rotor frame rotate together. Elastomeric seals275may provide hydraulic sealing at the axially translatable interface of the rotor frame251to the cylindrical base273. Hydraulic fluid may be communicated to and from the hydraulic fluid chamber270via passages277and a channel279through the motor output shaft225. Volume changes in the hydraulic fluid chamber270correspond to axial translation of the rotor frame251. Inner disc springs281between the motor output shaft225and the rotor frame251may provide a bias force when compressed, thus urging the rotor frame251away from the stator220S. Similarly, outer disc springs283between the bearing inner ring240B (or inner ring mounting hardware) and the rotor frame251may provide a bias force when compressed, thus urging the rotor frame toward the stator220S. The hydraulic system of the embodiment ofFIGS.2A and2Bmay be referred to as a linear hydraulic actuator system wherein the hydraulic actuator working surface area may correspond to the total rotor frame surface area opposing the vein plates271within the hydraulic fluid chambers270.

In the embodiment ofFIGS.2A and2B, the stroke travel of the linear hydraulic actuator corresponds one-to-one to the axial translation of the rotor frame251. Thus, small air gap changes (e.g., 1 mm) may advantageously be accomplished as rapidly and such linear hydraulic actuator arrangement may be advantageous in a two-state air gap control including a narrow air gap (e.g., 1 mm) at a first axial travel limit of the rotor frame251and a wide air gap (e.g., 2 mm) at a second travel limit of the rotor frame251.

In operation, pressurized hydraulic fluid may be supplied to the hydraulic fluid chamber270via the passages277and a channel279in the motor output shaft225to translate the rotor220R axially along the rotational axis (A) of the rotor220R away from the stator220S to increase the air gap260. The hydraulic fluid may be exhausted from the hydraulic fluid chamber270via the passages277and the channel279in the motor output shaft225to translate the rotor220R axially along the rotational axis (A) of the rotor220R toward the stator220S to decrease the air gap260. In one embodiment, translation of the rotor220R to decrease the air gap260may be effected by the force exerted by the compressed outer disc springs283between the bearing inner ring240B (or inner ring mounting hardware) and the rotor frame251and the magnetic attractive force of the permanent magnets255.

In the embodiment ofFIGS.3A and3B, the rotor220R may include at least one hydraulic fluid chamber270within the rotor frame251. Additional reference is made toFIGS.4A and4Bwhich schematically illustrate partial sectional views taken along the respective lines4A-4A and4B-4B inFIGS.3A and3B, respectively.FIGS.3A and3Billustrate one hydraulic actuator in section at one angular orientation of the rotor220R, it being understood that multiple such hydraulic actuators may be angularly distributed around the entire rotor220R for system balance requirements and overall hydraulic force multiplication proportional to the number of hydraulic actuators and hydraulic fluid chambers which may advantageously allow for lower system hydraulic pressures.FIGS.4A and4B, for example, illustrate two angularly balanced (i.e., 180 degrees separation) hydraulic fluid chambers270within the rotor frame251. Each hydraulic fluid chamber270may include a cavity272formed in the rotor frame251and a vein plate271extending radially outward from the motor output shaft225. The vein plate271is affixed to the motor output shaft225such that it is rotationally immovable relative to the motor output shaft225but is axially translatable, for example by a splined coupling between the motor output shaft225and the vein plate271. Thus, the vein plate271of the hydraulic fluid chamber is rotationally static with respect to the motor output shaft and rotates therewith and may translate axially along the rotational axis (A) of the rotor220R. The cavity272in the rotor frame251encloses the vein plate271and together therewith defines the hydraulic fluid chamber270. The rotor frame251is affixed to the motor output shaft225such that it is rotatable relative to the vein plate271and the motor output shaft225in accordance with volume changes of the hydraulic fluid chamber and is axially translatable with the vein plate271relative to the motor output shaft225. The rotor frame251axial and rotational movement is constrained to the motor output shaft by a lead screw coupling therebetween. In the illustrated embodiment ofFIGS.3A and3B, the lead screw coupling is shown as a low friction ball screw290. The ball screw290may include a ball return through the rotor frame251(not illustrated) or may be free-wheeling with a ball cage (not shown). Volume changes in the hydraulic fluid chamber270correspond to rotation of the rotor frame251relative to the vein plate271and the motor output shaft225which results in axial translation of the vein plate271and the rotor frame251via the ball screw290proportionally to the rotation of the rotor frame251corresponding to the screw pitch291of the ball screw290. Thus, the vein plate271and the rotor frame251may axially move together in accordance with volume changes in the hydraulic fluid chamber270(i.e., rotation of the rotor frame251relative to the vein plate271). Otherwise, the vein plate271and the rotor frame may be hydraulically locked together thus locking them axially and rotationally to the motor output shaft225. Elastomeric seals275may provide hydraulic sealing at the axially translatable interfaces of the rotor frame251to the motor output shaft225and of the vein plate271to the motor output shaft225. Hydraulic fluid may be communicated to and from the hydraulic fluid chamber270via passages277and a channel279through the motor output shaft225. Inner disc springs281between the motor output shaft225and the rotor frame251may provide a bias force when compressed, thus urging the rotor frame251away from the stator220S. Similarly, outer disc springs283between the bearing inner ring240B (or inner ring mounting hardware) and the rotor frame251may provide a bias force when compressed, thus urging the rotor frame toward the stator220S. The hydraulic system of the embodiment ofFIGS.3A,3B,4A and4Bmay be referred to as a rotary hydraulic actuator system wherein the hydraulic actuator working surface area may correspond to the total rotor frame surface area opposing the vein plate271within the hydraulic fluid chamber270.

In the embodimentFIGS.3A and3B, the stroke travel of the rotary hydraulic actuator corresponds to proportional fractional axial translation of the rotor frame251. For example, a pitch291of 2 mm on the ball screw290would result in 1 mm of axial translation of the rotor frame for 180 degrees of rotation thereof. Thus, small air gap changes (e.g., 1 mm) may advantageously be accomplishes in a substantially continuous manner and such a rotary hydraulic actuator arrangement may be advantageous in a continuous air gap control between a narrow air gap (e.g., 1 mm) at a first axial travel limit of the rotor frame251and a wide air gap (e.g., 2 mm) at a second travel limit of the rotor frame251.

In operation, pressurized hydraulic fluid may be supplied to the hydraulic fluid chamber270via the passages277and a channel279in the motor output shaft225to rotate the rotor frame251relative to the vein plate271and the motor output shaft225thereby translating the vein plate271and the rotor frame251via the ball screw290axially along the rotational axis (A) of the rotor220R away from the stator220S to increase the air gap260. The hydraulic fluid may be exhausted from the hydraulic fluid chamber270via the passages277and the channel279in the motor output shaft225to rotate the rotor frame251relative to the vein plate271and the motor output shaft225thereby translating the vein plate271and the rotor frame251via the ball screw290axially along the rotational axis (A) of the rotor220R toward the stator220S to decrease the air gap260. In one embodiment, translation of the rotor220R to decrease the air gap260may be effected by the force exerted by the compressed outer disc springs283between the bearing inner ring240B (or inner ring mounting hardware) and the rotor frame251and the magnetic attractive force of the permanent magnets255.

It is appreciated that in both embodiments ofFIGS.2A/2B andFIGS.3A/3B all axial translation forces of the rotor structure are advantageously contained withing the rotating components of the rotor and within the envelope generally between the axially static bearings240. Thus, while the bearing inner ring240B and/or the associated inner ring mounting hardware may be axially loaded through the outer disc springs283, the loading is ultimately transferred to the motor output shaft225and does not cross the bearing through the rolling elements.

FIGS.5A and5Bschematically illustrate a control mechanism for actuated and unactuated hydraulic control states, respectively, for establishing the rotor220R axial positions corresponding to narrow and wide air gaps260A and260B, respectively. Pressurized hydraulic fluid may be provided to the hydraulic fluid chamber270(FIG.5A) or hydraulic fluid may be exhausted from the hydraulic fluid chamber270(FIG.5B). A binary fluid control valve (e.g., solenoid controlled spool valve)501is illustrated in a first state (FIG.5A) connecting a pressurized hydraulic fluid from a high pressure supply side (S) of a hydraulic system503(e.g., pressurized line, pump outlet, accumulator, etc.) to the hydraulic fluid chamber270. The binary fluid control valve501is illustrated in a second state (FIG.5B) connecting the hydraulic fluid chamber270to a low pressure exhaust side (E) of the hydraulic system503(e.g., pump inlet, reservoir, sump, etc.).

Hydraulic pressure requirements may be determined in accordance with a force balance relationship in the case of the linear hydraulic actuator embodiment ofFIGS.2A and2Bmodeled by a force balance relationship as follows:

P×A=F[1]wherein P is the minimum hydraulic pressure required to move the hydraulic actuator;A is total hydraulic actuator working surface area; andF is force opposing the hydraulic actuator.
The force F opposing the hydraulic actuator may be primarily from permanent magnet attractive forces. The total hydraulic actuator working surface area A may be the summation of the hydraulic actuator working surface areas of all hydraulic fluid chambers.

In the embodiment ofFIGS.2A and2B, there are two hydraulic fluid chambers each having a respective hydraulic actuator working surface area, A1and A2wherein A=(A1+A2). Such relationship may be solved for the minimum hydraulic pressure P required to move the hydraulic actuator. An exemplary motor220may have permanent magnet attractive force of substantially 10,000 N and hydraulic actuator working surface areas of A1=A2=5890 mm2=0.00589 m2for a total hydraulic actuator surface working area A=A1+A2=2×0.00589 m2=0.01178 m2. Thus, the minimum hydraulic pressure required to move the hydraulic actuator P=10,000 N/0.01178 m2=848,896 N/m2(approx. 8.5 bar). The above relationship ignores bias force from the inner disc spring281between the motor output shaft225and the rotor frame251. Assuming a compressed spring force of substantially one-half the permanent magnet attractive force, inner disc spring281may reduce the force opposing the hydraulic actuator to substantially 5,000 N and the pressure required to move the hydraulic actuator P=5,000 N/0.01178 m2=424,448 N/m2(approx. 4.2 bar). It is thus appreciated that increasing total hydraulic actuator working surface area A and/or reducing force F opposing the hydraulic actuator may reduce the minimum hydraulic fluid pressure required to move the hydraulic actuator P. Total hydraulic actuator working surface area may be increased by axially stacking hydraulic fluid chambers (e.g., as illustrated in the dual chamber arrangements ofFIGS.2A and2B) and/or configuring multiple such arrangements angularly distributed around the entire rotor220R.

Hydraulic pressure requirements may be determined in accordance with a torque balance relationship in the case of the rotary hydraulic actuator embodiment ofFIGS.3A and3Bmodeled by a torque balance relationship as follows:

P×(A×r)=T[2]wherein P is the minimum hydraulic pressure required to move the hydraulic actuator;A is total hydraulic actuator working surface area;r is the equivalent radius of the hydraulic actuator working surface; andT is torque opposing the hydraulic actuator.
The force T opposing the hydraulic actuator may be primarily from permanent magnet attractive forces. The total hydraulic actuator working surface area A may be the summation of the hydraulic actuator working surface areas of all hydraulic fluid chambers.

In the embodiment ofFIGS.3A,3B,4A and4Bthere are two chambers each having a respective hydraulic actuator working surface area, A1and A2wherein A=(A1+A2). Such relationship may be solved for the minimum hydraulic pressure P required to move the hydraulic actuator. An exemplary motor220may have permanent magnet attractive force of substantially 10,000 N. This force may be related to the torque opposing the hydraulic actuator T through a ball screw relationship as follows:

wherein F is the permanent magnet attractive force; andL is the axial distance per revolution (i.e., lead).
Thus, in the present example, the torque opposing the hydraulic actuator T=(10,000 N×2 mm)/2π=(10,000 N×0.000002 m)/2π=0.0032 N-m.

The exemplary motor220may have hydraulic actuator working surface areas of A1=A2=600 mm2=0.0006 m2for a total hydraulic actuator surface working area A=(A1+A2)=2×0.0006 m2=0.0012 m2. The exemplary motor220may have a hydraulic actuator working surface inner radius R1=25 mm and outer radius R2=55 mm. Thus, the equivalent radius r of the hydraulic actuator working surface may be determined from the following relationship:

Thus, in the present example, r=√[(25 mm)2+(55 mm)2]/2=0.000030 m and the minimum hydraulic pressure required to move the hydraulic actuator P=T/(A×r)=0.0032 N-m/(0.0012 m2×0.000030 m)=88,889 N/m2(approx. 0.89 bar). The above relationship ignores bias force from the inner disc spring281between the motor output shaft225and the rotor frame251. Assuming a compressed spring force of substantially one-half the permanent magnet attractive force, inner disc spring281may reduce the force opposing the hydraulic actuator to substantially 5,000 N and the pressure required to move the hydraulic actuator P by one half or about 44,445 N/m2(approx. 0.44 bar). It is thus appreciated that increasing total hydraulic actuator working surface area A and/or reducing force F opposing the hydraulic actuator may reduce the minimum hydraulic fluid pressure required to move the hydraulic actuator P. Total hydraulic actuator working surface area may be increased by adding hydraulic fluid chambers angularly distributed around the entire rotor220R. However, it is understood that more hydraulic fluid chambers on the same axial location may further limit the angular range of motion of the hydraulic actuators.

FIG.6illustrates exemplary torque (TQ) versus motor speed (S) curves corresponding to motor operation at relatively narrow air gaps260A (i.e.,FIGS.2A and3A) and relatively wide air gaps260B (i.e.,FIGS.2B and3B). Relative motor torque (TQ) is plotted along the vertical axis with increasing torque away from zero at the origin (O) and relative motor speed (S) is plotted along the horizontal axis with increasing speed away from the origin (O). Narrow air gap260A operation may exhibit a constant torque output TQ1from zero speed until a base speed B1corresponding to the narrow air gap260A, whereafter motor operation in the constant power region at higher speeds exhibits characteristic reduced torque output until a terminal speed S1. Wide air gap260B operation may exhibit a constant torque output TQ2from zero speed until a base speed B2corresponding to the wide air gap260B, whereafter motor operation in the constant power region at higher speeds exhibits characteristic reduced torque output until a terminal speed S2. Thus, it is appreciated that the narrow air gap260A may have a higher available torque at lower motor speeds and more limited torque output at higher motor speeds. It is further appreciated that the wide air gap260B may have a lower available torque at lower motor speeds but less limited torque output at higher motor speeds. The extended speed and torque operating range of the motor with a wide air gap260B may be attributed to the field weaking achieved by the larger physical air gap.

In an automotive application such as an electric propulsion system101in a vehicle100as described herein, it is appreciated that adjustable air gaps in an axial flux motor220may provide for selective, mechanical flux weakening. Such selective control may be effected by way of the binary fluid control valve501hydraulic system503as described herein with respect toFIGS.5A and5B. The binary fluid control valve may be501may be controlled in response to a control signal from a control module, for example the vehicle controller103and/or the motor controller105as described herein with respect toFIG.1. Advantageously, mechanical flux weakening through widening the air gap above the motor base speed corresponding to a narrow air gap may reduce copper losses associated with alternative electrically induced field weakening controls (e.g., current injections). Advantageously, the air gap may be controlled wider at low vehicle loads to reduce motor cogging torque. Efficiency gains may be realized by operating wider air gaps when stator core losses exceed copper losses. And, widening the air gap may advantageously extend useful torque/speed operation in the constant power operating region of the motor above the base speed corresponding to a narrow air gap through the beneficial mechanical flux weakening. It is further appreciated that in an automotive application such as an electric propulsion system101in a vehicle100as described herein, adjustable air gaps in an axial flux motor220may provide for selective operator selectable vehicle response mode control. For example, an economy mode associated with light acceleration and steady powertrain loading may benefit from a wide air gap setting. Similarly, a sport mode associated with spirited acceleration, busy throttle use and more rapid throttle tip-ins and tip-outs may benefit from a narrow air gap at least with the constant torque operating region below the motor base speed corresponding to a narrow air gap.

All numeric values herein are assumed to be modified by the term “about” whether or not explicitly indicated. For the purposes of the present disclosure, ranges may be expressed as from “about” one particular value to “about” another particular value. The term “about” generally refers to a range of numeric values that one of skill in the art would consider equivalent to the recited numeric value, having the same function or result, or reasonably within manufacturing tolerances of the recited numeric value generally. Similarly, numeric values set forth herein are by way of non-limiting example and may be nominal values, it being understood that actual values may vary from nominal values in accordance with environment, design and manufacturing tolerance, age and other factors.

When an element such as a layer, film, region, or substrate is referred to as being “on” another element, it can be directly on the other element or intervening elements may also be present. In contrast, when an element is referred to as being “directly on” another element, there are no intervening elements present. Therefore, unless explicitly described as being “direct,” when a relationship between first and second elements is described in the above disclosure, that relationship may be a direct relationship where no other intervening elements are present between the first and second elements but may also be an indirect relationship where one or more intervening elements are present (either spatially or functionally) between the first and second elements.

One or more steps within a method may be executed in different order (or concurrently) without altering the principles of the present disclosure. Further, although each of the embodiments is described above as having certain features, any one or more of those features described with respect to any embodiment of the disclosure can be implemented in and/or combined with features of any of the other embodiments, even if that combination is not explicitly described. In other words, the described embodiments are not mutually exclusive, and permutations of one or more embodiments with one another remain within the scope of this disclosure.