EP1110008A1 - Regenerative adaptive fluid control - Google Patents

Regenerative adaptive fluid control

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Publication number
EP1110008A1
EP1110008A1 EP98931246A EP98931246A EP1110008A1 EP 1110008 A1 EP1110008 A1 EP 1110008A1 EP 98931246 A EP98931246 A EP 98931246A EP 98931246 A EP98931246 A EP 98931246A EP 1110008 A1 EP1110008 A1 EP 1110008A1
Authority
EP
European Patent Office
Prior art keywords
fluid
energy
motor
load
feedback control
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Withdrawn
Application number
EP98931246A
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German (de)
French (fr)
Inventor
Robert M. Lisniansky
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Individual
Original Assignee
Individual
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Publication date
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Publication of EP1110008A1 publication Critical patent/EP1110008A1/en
Withdrawn legal-status Critical Current

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B21/00—Common features of fluid actuator systems; Fluid-pressure actuator systems or details thereof, not covered by any other group of this subclass
    • F15B21/14—Energy-recuperation means
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B1/00—Installations or systems with accumulators; Supply reservoir or sump assemblies
    • F15B1/02—Installations or systems with accumulators
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B2203/00—Motor parameters
    • F04B2203/12—Motor parameters of rotating hydraulic motors
    • F04B2203/1201—Rotational speed

Definitions

  • the present invention relates primarily to a fluid motor position feedback control system, such as the electrohydraulic or hydromechanical position feedback control system, which includes a fluid motor, a primary variable displacement pump, and a spool type directional control valve being interposed between the motor and the pump and being modulated by a motor position feedback signal. More generally, this invention relates to the respective fluid motor output feedback control systems and to the respective fluid motor open-loop control systems. In a way of possible applications, this invention relates, in particular, to the hydraulic presses and the motor vehicles.
  • T h e hydrauJ lc Cluid motor is usually driving a variable load.
  • the exhaust and supply fluid pressure drops across the directional control valve are c h anged, which destroys the linearity of a static speed characteristic describing the fluid motor speed versus the valve spool displacement.
  • Al a r ⁇ ault, a system fain and tht related qualities, such as tht dynamic performance and accuracy, are all tht function* of tht variable load.
  • an energy efficiency of tht petition feedback control is also a function of the variable load.
  • a hydraulic press is an impressive example of the heavy loaded hydraulic motor-mechanism. Tht load conditions are changed substantially within each press circle, including approaching the work, compressing tht fluid, tht working stroke , d ⁇ compretaing tht fluid, and tht rtturn stroke.
  • the first typical hydraulic schematic includes a three- ay directional control valve In combination with the two counteractive ( ⁇ xpansibl ⁇ j chambers.
  • the first of these chambers is controlled by said three-way valve «hich is alto oonntcttd to the pressure and tank lines of the fluid power means.
  • the second chamber Is under a relatively constant pressure provided by said pressure lint.
  • the second typical schematic includes a four-way directional control valve in combination with the two pount ⁇ ractive chambers. Both of thtst chambers are controlled by the four— way valve which is also connected to the pressure and tank lints of tht fluid power means.
  • this schematic it is not possible to automatically maintain an exhaust fluid across the four-way valve without encountering which can also' be viewed as a schematic operation interference with the position feedback control system.
  • a chamber's pressure signal which Is needed for maintaining the exhaust fluid pressure drop, must be switched over from one chamber to the other in exact accordance with a valve spool transition through a neutral spool position, where the chamber lines are switched over, to avoid damaging the spool valve flow characteristics.
  • a pressure differential between the o chambers at the neutral spool position *U1 affect the pressure drop regulation and may generate the dynamic unstability o the position feedback ccntrol system.
  • the structural weakness of the conventional fluid motor position feedback control systems can be still further characterized by that these systems are not equiped for regenerating a load related energy, such as a kinetic energy of a load mass or a compressed fluid energy of the fluid motor-cylinder. As a result, this load related energy is normally lost.
  • load related energy such as a kinetic energy of a load mass or a compressed fluid energy of the fluid motor-cylinder.
  • the present invention is primarily aimed to improve the ⁇ qualities and energy efficiency of the fluid motor position feedback control system, such as the el ⁇ ctrohydraulic or hydromechanical position feedback control system, operating usually in the variable load environments.
  • the implementation of these interrelated steps and conditions is a way of transition from the conventional fluid motor position feedback control systems to the load adaptive fluid motor position feedback control systems.
  • These load adaptive systems can generally be classified by the amount of controlled and loadable chambers of the fluid motor, by the spool valve design configurations, and by the actual shape of the spool valve flow characteristics. In a case hen only one of two counteractive chambers cf the fluid motor is controllable, the fluid motor can be loaded only n one direction.
  • the controlled chamber Is connected to the three-way spool valve which also has a supply pow «r line and an exhaust power line. In this case, the second chamber is under a relatively constant pressure su plied by an independent source of fluid power.
  • the fluid motor can be loaded in only one or in both directions.
  • the controlled chambers are connected to a fiv ⁇ - * .ay spool valve which also has a common supply power line and two separate exhaust power lines.
  • * ' hen the fluid motor is loaded in only one direction, 'only one of two exhaust lines is aJLso a cbunt ⁇ rpr ⁇ ssure line.
  • n ' hen the fluid motor is loaded in both directions, both exhauat lln ⁇ are used as oount ⁇ rpr ⁇ ssur ⁇ lines.
  • each count ⁇ rpressur ⁇ line is provided *ith an exhaust line pressure drop regulator.
  • l*hich is modulated by an exhaust line pressure drop feedback signal which is measured between this counterpressure line and the r»l.atji,d_ chamber.
  • a supply fluid flo* rate is being monitored continuously by the primary variale displacement pump of the fluid power means. Maintaining the supply fluid pressure drop is also a way of regulating the hydraulic power delivered to the spool type directional control valve.
  • the exhaust fluid pressure i rz p across the spool valve all the flow is being released from the counterpressure line through the exhaust line pressure drop regulator to the tank. Counterpressure may be created in the counterpressure line only for a short time while the hydraulic fluid in the preloaded chamber is being decompressed. However, .the control over the decompression is critically Important for improving the system's dynamic performance potential.
  • a family of load adaptive fluid position servomechanisms may include the three-, four-, five-, and six-way directional valves
  • the three-way spool valve is used to provide the individual pressure and counterpressure lines for only one controllable chamber.
  • the six-way spool valve is used to provide the separate supply and exhaust lines for each of two controllable chambers.
  • the five-way spool valve can be derived from the six-way spool valve by connecting together two separate supply lines.
  • the four-way spool valve can be derived from the five-*ay spool valve by connecting together -two separate exhaust lines.
  • the four-way spool valve does create a problem of schematic operation interference between the position feedback control and the regulation of pressure drops, as it is already explained above.
  • the fluid motor is provided with at least one controlled and loadable chamber, and that this chamber is provided with the pressure-compensated spool valve flow characteristics.
  • These pressure-compensated flow characteristics are shaped by the related exhaust line pressure drop feedback control system which includes the exhaust line pressure drop regulator and by the related supply line pressure drop feedback control system which Includes the primary variable displacement pump.
  • the deqired (linear or unlinear) shape of the spool valve flow characteristics is actually implemented by programming the supply and exhaust line pressure drop command signals of the supply and exhaust line pressure drop feedback control systems, respectively. Some possible principals of programming these command signals are illustrated below.
  • the supply and exhaust line pressure drop command signals are set approximately constant for linearizing the pressure-compensated spool valve flow characteristics.
  • the related adaptive hydraulic (electrohydraulic or hydromechanical) position servo echanismsc can be referred to as the linear adaptive servomechanisms, or as the fully-compensated adaptive servomechanisms. Still other method of programming the piessure drop command signals can be specified with respect to the linear adaptive servomechanisms, as it is illustrated below- -by points 2 to 5.
  • a load related energy such as a kinetic energy of a load mass or a compressed fluid energy of the fluid motor-cylinder.
  • I is still further object of this inventiion to develop a concept of load adaptive exchange of energy between the fluid motor and load means and the energy accumulating means of the load adaptive energy regenerating system.
  • the load adaptive regeneration of the load related energy of the fluid motor and load means can be viewed as a part (or as a larger part) of a complete circle of the load adaptive exchange of energy between the fluid motor and load means and the energy accumulating means.
  • the regenerative adaptive fluid control makes it possible to combine the load adaptive primary power supply and the load adaptive regeneration of energy for maximizing the over-all energy efficiency and performance potentials of the fluid motor control systems.
  • 1 is still further object of, this invention to develop the high energy-ef icient, load adaptive hydraulic presses utilising the regenerative adaptive fluid control.
  • Fig.l sho th ⁇ adaptive fi uid 3 ⁇ r vomecha*i 3m having only one controllable chamber.
  • Pig.2 shows a power supply schematic version.
  • Fig-3- ⁇ is a generalization of Fig.l.
  • Fig. -D illustrates the flow characteristics of valve 2.
  • Pi ⁇ shows the adaptive fluid s ⁇ rvom ⁇ chanlsm having two controllable chambers but loadable only in one direction
  • Fig.5- ⁇ is a generalization of Fig. .
  • Fig.5-B Illustrates the flow characteristics of valve 2.
  • Fig.6 shows th ⁇ adaptive fluid s ⁇ rvomechanism having two controllable chambers and loadable in both directions.
  • Flg.7- ⁇ is a generalization of Fig.6.
  • Flg.7-n illustrates the flow characteristics of valve Z .
  • rig. ⁇ shows a generalized model of adaptive fluid position servomechanisms .
  • Fig. illustrates the concept of load adaptive regeneration of energy.
  • Fig.10 shows the adaptive fluid servomechanism having
  • Fig.11 shows the adaptive fluid servomechanism having an independent energy regenerating circuitry.
  • Flp..l2 is a modification of Fig.H for the hydraulic press type applications.
  • Fig.13 shows a generalized model of the regenerative adaptive fluid motor output feedback control systems.
  • Fig.l' . shows a generalized model of the regenerative adaptive fluid motor velocity feedback control systems.
  • Fig.15 shows a generalized model of the regenerative adaptive fluid motor open-loop control systems.
  • Fig.i ⁇ is a modification of Fig.11 for the .motor vfehicle type applications.
  • Fig.17 shows a regenerative adaptive drive system for the motor vehicle type applications.
  • Fig. IB shows a regenerative adaptive drive system having a hydraulic accumulator.
  • s h ows a regenerative adaptive drive system having the combined energy , regenerative means.
  • F ig. 20 s h ows a regenerative adaptive drive system having a variable displacement motor driving the load.
  • I r ig.2) shows a regenerative adaptive drive system having a regenerative braking pump.
  • Fig.22 shows a modified regenerative system having a hydraulic accumulator.
  • Fig.23 shows the load adaptive displacement means of the assisting supply line pressure drop feedback control system.
  • Fig.24 shows the load adaptive displacement means of the energy recupturing pressure drop feedback control system.
  • Fig.25 illustrates a stop-arid-go energy regenerating circle.
  • Fig.26 shows a modified regenerative system having the combined energy regenerating means.
  • Fig.27 shows a generalized regenerative system having a built-in regenerating circuitry.
  • Fig.28 shows a regenerative adaptive drive system having a supplementary output motor.
  • Fig.29 shows a generalized regenerative system having a supplementary output motor. DESCRIPTION OF THE INVENTION
  • Fig.l shows a simplified schematic of the load adaptive fluid motor position feedback control system having only one controllable chamber.
  • the moving part 21 of the fluid motor-cylinder 1 is driven by two counteractive expansible chambers - chambers 10 and 11, only one of which - chamber ID - is controllable and can be loaded.
  • the second chamber - chamber 11 - is under a relatively low (and constant ) pressure P Q supplied by an independent pressure source.
  • This schematic is developed primarily for the hydraulic press type applications. ⁇ s it is already mentioned above, the load conditions are changed substantially within each press circle including approaching the work, compressing the fluid (in chamber 10), the working stro k, decompressing the fluid (in chamber 10), and the return strock.
  • the schematic of Fig.l further includes the hydraulic power supply means 3-1 having a primary variable displacement pump powering the pressure line 51.
  • the three-way spool-type directional control valve 2 is provided with three hydraulic power lines Including a motor line — line LI - connected to line 15 of cha ⁇ b ⁇ r 10 , th ⁇ supply power line L2 connected to pressure line 51 . and the exhaust power line LJ. Lines L2 and IQ are commutat ⁇ d *lth line LI by th ⁇ spool valve 2. To consider all the picture, Fig.l should be studied together with the relate -supplementary figures Z , 3- ⁇ , and 3-13.
  • Th ⁇ block represents a generalized model of the optional position feedback control means. This block is needed to actually make-up th ⁇ fluid motor position feedback control .system, which is capable of regulating the motor position X, of motor 1 by employing the motor position feedback signal CX ⁇ , where coefficient "C" is, usually, constant. Th ⁇ motor position feedback signal CX, is generated -by a motor position sensor, which is included into block '» ⁇ and is connected to the moving part 21 of th ⁇ hydraulic fluid motor 1.
  • An original position feedback control error signal ⁇ X Qr is produced aa a difference between the position input-command signal X and the motor position feedback signal CX ⁇ .
  • There are at least two typical fluid motor position feedback control systems the electrohydraulic and hydromechanical position feedback control systems.
  • th ⁇ equation —CX,l o the like is simulated by electrical means located within block .
  • the block may also include the electrical and hydraulic amplifiers, an electrical torque motor, the stabilization— — optimization technique and other components to properly amplify and condition said signal ⁇ X Qr for modulating said valve 2.
  • theVposition feedback control error signal ⁇ x is derived in accordance with a difference between the position input-command signal X and the output position signal X..
  • Th ⁇ exhaust line pressure drop regulator 3-3 is introduced to make up the exhaust line pressure drop feedback control system which is capable of regulating the exhaust fluid pressure drop across valve 2 by varying th ⁇ counterpressure rate P., in th ⁇ exhaust po-v ⁇ r lin ⁇ 'L3.
  • This exhaust fluid pressure drop is represented by th ⁇ exhaust line pressure drop feedback signal, *hich is equal ? Q - > — P3 and is measured between th ⁇ exhaust power line L3 and th ⁇ related exhaust signal line SL3 connected to line LI.
  • Th ⁇ regulator 3-3 is connected to th ⁇ exhaust power line L3 and to th ⁇ tank line 52 and is modulated by an exhaust line pressure drop feedback control error signal, which is produced in accordance with a difference between th ⁇ exhaust line pressure drop command signal ⁇ .
  • the primary variable displacement pump of fluid power supply means 3-1 ( pump 59 on Pig.2 ) is introduced to make-up th ⁇ supply line pressure drop feedback control system, which is capable of regulating the supply fluid pressure drop across valve 2 by varying th ⁇ pressure rate P 2 in the supply power line L2 by varying th ⁇ supply fluid flow rate in said line L2 by said variable displacement pump.
  • This supply fluid pres ⁇ sure drop is represented by the supply line pressure drop feedback signal, which is equal P 2 —' P Q2 *r ⁇ is measured between line L2 (through line 32 on Pig.2) and th ⁇ related s ⁇ pSiX line 3L2 connected to lint LI.
  • a variable delivery means 56 of pump 58 is modulated by a supply line pressure drop feedback control error signal, which is produced in accordance with a difference between th ⁇ supply line pressure drop command signal ⁇ P 2 *nd the supply line pressure drop feedback • signal ? 2 — P 02 •
  • Th ⁇ pressure maintained in th ⁇ supply power line L2 by th ⁇ supply line pressure drop feedback control system is - ⁇ P 2 *n d can be Just slightly above what is require for chamber 10 to overcome the load.
  • the counterpressure maintained in the exhaust power line L3 by the exhaust line pressure drop feedback control system is P 03 """" ⁇ P3 *nd can be Just slightly below the pressure P Q - —r P Q2 in chamber 10 .
  • the pressure drop command signals ⁇ P 2 and ⁇ P • the pressure P Q and their interrelationship are selected for linearising th ⁇ spool valve flow characteristic ( PSBK, ⁇ X ) without "running a risk” of full decompressing the hydraulic motor ( chamber 10 ) and generating th ⁇ hydraulic shocks in the hydraulic system.
  • the pressure P Q has to compress the hydraulic fluid in chamber 10 to such an extent as to prevent the full decompression under the dynamic operation conditions. In the absence of static and dynamic loading, the pressure P 10 in chamber 10 is fixed by th ⁇ pressure P Q applied to l o
  • Th ⁇ regulator 3-3 is opened by a force of th ⁇ spring shown on Fig.l and is being closed to provide th ⁇ counterpressure P- j only after th ⁇ actual pressure drop P Q , — p_ exceeds its pr ⁇ install ⁇ d value ⁇ . which is defined by the spring force Practically, at th ⁇ very beginning of th ⁇ return stroke, *h ⁇ n the regulator 3-3 has to enter into th ⁇ operation, th ⁇ controllable chamber 10 is still under th ⁇ pr ⁇ ur ⁇ . It means that regulator 3-3 is preliminarily closed and is ready to provide th ⁇ count ⁇ rpr ⁇ sure P. , which is being maintained by regulator 3-3 only for a short time of decompressing chamber 10.
  • the schematic of Fig. 2 is a disclosure of block 3-1 shown on Fig.l.
  • This schematic includes the primary variable displacement pump 58, which is connected through line 30 and check valve 44 to the pressure line 51.
  • a relatively low pressure, high capacity fluid power supply 50 (such as a centrifugal pump) is also connected through line 5- and check valve O to the pressure line ji.
  • Th ⁇ primary motors (such as electrical motors driving the pumps are not sho- on Pig.2.
  • Th ⁇ variable delivery means 56 of pump 58 includes a variable displacement mechanism of this pump.
  • the tank lines 52 and 36 are collected by the oil tank 62.
  • the pressure line 51 can be protected by the maximum pressure relief valve *hich is not sho*n on Pig.2. Th ⁇ maximum pressure in line 5 can also be restricted by using the variable delivery means 5 of pump 58. In general, the maximum pressure relief valves can also be used to protect other hydraulic lines.
  • a relatively low pressure fluid from th ⁇ high capacity fluid power supply 50 is introduced through check valve ⁇ 0 into th ⁇ pressure line 51 to increase th ⁇ speed limit of th ⁇ hydraulic cylinder 1 ( Fig.l ), as th ⁇ pressure rate in line 1 is sufficiently declined.
  • th ⁇ hydraulic po/.er supply 50 is being entered into th ⁇ operation Just after th ⁇ spool of valve 2 passes its critical point, beyond vhich the pressure P- in line 51 is dropped b ⁇ lo* the minimum regulated pressure P 2min '
  • Th ⁇ schematic shown on Fig.l is a sy m ⁇ trical, relative to the chambers 10 and 11.
  • the functional operation of this schematic can be still better visualized by considering its generalized model, hich is presented on ?ig.3-A and is accompanied by th ⁇ related pressure-compensated flow characteristic P- ⁇ K- ⁇ ⁇ x of valve 2.
  • Th ⁇ fluid power means 3 shown on Fig.3-A combine the fluid power supply means 3-1 and the regulator.* 3-3 , which are shown on Fig.l- The concept of preventing a substantial schematic operation interference.
  • Fig. shows a simplified schematic of the load adaptive fluid motor position feedback control system having two controllable chambers but loadable only in one direction.
  • This schematic is also developed primarily for the hydraulic press type applications, is provided with the five-way spool valve 2, and is easily understood when compared with Fig.l.
  • the line 12 of chamber 11 is connected to line L4 ol valve 2.
  • the loadable chamber 10 is controlled as before.
  • the chamber 11 is commutated by valve 2 with the supply power line Lb and with the "unregulated" separate exhaust line L5.
  • the supply power line L6 is connected to line L2 but is also considered to be “unregulated", because the supply signal line SL2 is communicated (connected) only with chamber 10.
  • the exhaust line L5 is, in fact, the tank line.
  • equation (1) can be generalized as:
  • Th ⁇ pressures P 1Q and P ⁇ have to be hi ⁇ h enough to prevent th ⁇ full decompression of chambers 10 and 11 under th ⁇ dynamic operation conditions.
  • th ⁇ pressure drop command signals ⁇ 2 and ⁇ i have to be small enough to improve th ⁇ system energy efficiency.
  • Th ⁇ schematic shown on Fig is assymm ⁇ trical, relative to th ⁇ chambers 10 and 11. Th ⁇ functional operation of this schematic can be still better visualized by considering its generalized model, which is presented on Fig.5-A and is accompanied ' by th ⁇ related flow characteristics ⁇ I "" 1 ⁇ l ⁇ * and flow power means 3 shown on Fig.5-A, combine the fluid power supply means 3-1 and the regulator 3-3 » which are shown on Fig . 4 .
  • the schematic shown on Fig.6 is related to the load adaptive hydraulic position servomechanism having two controllable chambers and loadable in both directions.
  • This schematic is provided with th ⁇ five-way spool valve and is easily understood when compared with Fig. .
  • the loadable chamber 10 is controlled as before except that the supply signal line SL2 is communicated ( commutated ) with chamber 10 through. check valve 5.
  • Th ⁇ second loadable chamber — chamber 11 — is commutated by valve 2 with th ⁇ supply power line ,L6 and with th ⁇ exhaust power line L5.
  • Th ⁇ line L6 is connected to line L2.
  • Th ⁇ supply signal line SL2 is also communicated ( commutated ) with chamber 11 through check valve 6.
  • Th ⁇ exhaust line L5 is a separate counterpressure line which is provided with an additional exhaust line pressure drop feedback control system including an additional exhaust line pressure drop regulator 3-k which is shown on Fig.6..
  • ⁇ P ⁇ is th ⁇ related pressure drop command signal.
  • Th ⁇ check valve logic makes it possible for th ⁇ line SL2 to select one of two chambers, whichever has th ⁇ higher pressure rate, causing no problem for maitaining the supply fluid pressure drop across valve 2, as well as for the dynamic stability of the fluid motor position feedback control system.
  • a very small throttle valve 19 connecting line SL2 with the tank line 52, is helpfull in extracting signal P Q2 .
  • Th ⁇ schematic shown on Fig.6 is symmetrical, relative to th ⁇ chambers 10 and 11.
  • the motor load which is not shown on the previous schematics, is applied to the moving part 21 of th ⁇ hydraulic fluid motor 1.
  • This load is usually a variable load, In terms of its magnitude and ( or) direction, and may generally include the static and dynamic components.
  • Th ⁇ statio loading components are th ⁇ on ⁇ -dir ⁇ ctlonal or two-directional forces.
  • the dynamic (inertia) loading component is produced by accelerating and decelerating a load mags ( i nc i u di n the mass of moving part 21) and is usually a two-directional force. If th ⁇ fluid motor 1 is loaded mainly only in one direction by a static force, the schematic of Fig.l or Fig.l* is likely to be selected.
  • fluid motor 1 is provided with at least one controlled and loadable chamber, and that this chamber is provided with the pressure-compensated spool valve flow characteristics.
  • This idea can be best illustrated by a model of Fig.8 which is a generalization of Fig.3-A, Fig.5-A, and Fig.7-A.
  • the block 5 of Fig .8 combines fluid motor means (the fluid motor 1) and spool valve means (the spool valve 2), which are shown on previous schematics.
  • load adaptive fluid motor position feedback control systems being considered are not limited to the hydraulic press type applications.
  • the supply and exhausty lines L2, L3, L5, L6 are commutated with the chamber lines LI L4, the related signal lines SL2, SL3, SL5, SL6 must be communicated accordingly with th ⁇ same chamber lines 11, L ⁇ .
  • signal Un ⁇ s SL2, SL3, S15, SL6 with th ⁇ chambers can be provided by connecting or commutating these signal lines with th ⁇ chambers. Having th ⁇ separate supply and exhaust power lines for each controllable chamber, as well as having only one loadable chamber, makes it possible to eliminate the need for commutating these signal lines.
  • the commutation of supply signal line SL2 can be accomplished by such commutators as follows:
  • Fig.6 can be modified by replacing the first-named commutator by the second-named commutator.
  • the modified schematic is of a very general nature and is applicable to the complex load environments.
  • electrohydraulic position servomechanisms is quite similar to that of electromecha- nica ⁇ vg rvon ⁇ cTianisms. It is to say that in the case of electrohydraulic position servomechanisms, the electrical portion of block — including th ⁇ optional position sensor but excluding th ⁇ electrical torque motor — can also be characterize by the analogy with th ⁇ comparable portion of th ⁇ electric motor position feedback control systems — — see, for example, th ⁇ books already named above.
  • the motor position X ⁇ is the position of moving part 21 (piston, shaft and so on) of the fluid motor 1. in fact, the motor, position X ⁇ can also be viewed as a mechanical signal — the output position signal of the fluid motor position feedback control system being considered.
  • the motor position X j is measured by the position feedback control means due to the position sensor, which is included into block ⁇ and is connected to the moving part 21 of th ⁇ fluid motor 1.
  • a n electromechanical position sensor can be analog or digital.
  • the analog position sensor employs an analog transducer, such as a linear variable differential transformer, a synchro transformer, a resolver and so on.
  • the digital position sensor may include a digital transducer, such as an optical encorder.
  • the digital positionsensor can also be introduced by an analog-digital combination, such as the resolver and the resolver-to-digital converter — see, for example, chapter l of th ⁇ above named book of Analog devices, Inc. .
  • th ⁇ motor position feedback signal CX ( or th ⁇ like ) is generated by the electromechanical sensor in a form of th ⁇ electrical, analog or digital, signal, respectively. __,,.
  • th ⁇ position input - -command signal X is also th ⁇ electrical, analog or digital, signal, respectively.
  • Th ⁇ position input-command signal X can be generated by a variety of components— from a simple potentiometer to a computer.
  • th ⁇ mechanical position sensor is simply a mechanical connection to th ⁇ moving part 21 of th ⁇ fluid motor 1.
  • th ⁇ motor position feedback signal CX is a mechanical signal.
  • Th ⁇ , position input-command signal X is also a mechanical signal.
  • the spool of valve 2 is most often actuated through the hydraulic amplifier of th ⁇ position feedback control means.
  • the spool valve 2, the hydraulic amplifier, and the electrical torque motor are usually integrated into what is called an "electrohydraulic servovalve”.
  • the spool of valve 2 is also most often actuated through the hydraulic amplifier of the position feedback control means.
  • Th ⁇ spool valve 2 and th ⁇ hydraulic amplifier are usually integrated into what is called a "s ⁇ rvovalv ⁇ ? 10.
  • Still more comprehensive descriptioon of the optional position feedback control means ( block ) can be found in the prior art patents and publications including th ⁇ books already named above.
  • th ⁇ compressed hydraulic fluid is substantial in defining the system energy efficiency, a regeneration of this energy can ba Justified.
  • Pig * is originated by combining Fig.l and Fig.2.
  • th ⁇ regulator 3-3 is replaced by a variable displacement actor 6 having a variable displacement means 67 t a pressure line 77, and a tank line 7
  • Th ⁇ motor 65 is connected through line 77 to line L3 and has a •c ⁇ emwn shaft" 72 with th ⁇ variable displacement pump 58.
  • Th ⁇ variable displacement means 67 is modulated by th ⁇
  • Th ⁇ exhaust line pressure drdj contfol system Including motor 65, maintains the exhaust fluid pressure drop T Q -> — P, across spool valve 2 by varying th ⁇ counterpressure '3 ⁇ '03 —" ⁇ F3 in the exhaust line L3 by th ⁇ variable displacement means 67 .
  • a fly-wheel 9 is attached to the shaft 72 and is driven by motor 65.
  • the pump 58 is generally driven JyTN motor 100, by the motor 65 and by .
  • Fig.9 also shows the frame 190 ( of hydraulic press 192 ) against which the chamber 10 of cylinder 1 is loaded.
  • load adaptive regeneration of energy is further illustrated by considering the load adaptive, position feedback controlled, variable speed drive systems for the motor vehicle type applications (see figures 10 and 11), where a kinetic energy associated with a mass of the motor vehicle is substantial in defining the over-all energy efficiency. It will be shown that load adaptability of these efficient and flexible drive systemS j makes it easy to create the schematic conditions under which the energy accumulated during decelerating the motor vehicle is reused for accelerating the vehicle.
  • the load adaptive, position feedback controlled, variable speed drive systems may incorporate a built-in regenerating circuitry or an independent regenerating circuitry.
  • the drive system incorporating the built-in regenerating circuitry is shown on Fig.10 which is originated by combining Fig.6 and Fig.2.
  • the fluid power supply of Fig.2 is represented on Fig.10 mainly by pump 58.
  • the regulator 3-3 is not needed now and, therefore, is not shown on Fig.10.
  • the regulator 3-4 is replaced by a variable displacement motor 66 having a variable displacement means 68 , t ar. ⁇ line 7 , and pressure line 78 which is connected to line L5- Th ⁇ hydraulic cylinder 1 shown on Pig.
  • Th ⁇ fly-wheel 9 ⁇ is attached to th ⁇ common shaft 72 connecting pump 58, motor 66, and th ⁇ primary motor 100 of the motor vehicle.
  • Th ⁇ variable displacement means 68 is modulated by th ⁇ exhaust line pr ⁇ ⁇ aur ⁇ dr feedback signal, which is equal P 0 «— ? « * «* « ⁇ asu Hv-fJ
  • I, li n ⁇ ,5 (through line 76) and th ⁇ related signal line 3L5 « Th ⁇ exhaust line pressure drop feedback control system Including the variable displacement motor 66 , regulates the exhaust fluid pressure drop PQ* ⁇ — P « across spool valve 2 by varying th ⁇ counterpressure * « - * 0 « " ⁇ " ⁇ '5 *** the exhaust power line L5 by th ⁇ variable displacement means 68.
  • ene motor position command signal X being varied with the constant speed, will f ⁇ n ⁇ rat ⁇ a relatively constant velocity of motor 1 and the positional lag ⁇ l proportional to this velocity.
  • th ⁇ shaft v ⁇ looity of motor 1 can be controlled by the speed of varying th ⁇ motor position command signal X Q .
  • th ⁇ deceleration of th ⁇ motor vehicle th ⁇ kinetic energy accumulated by ' a mass of th ⁇ motor vehicle (load 96) is transmitted through motor 66 to th ⁇ fly-wheel 9 ⁇ .
  • Fig.11 The load adaptive, position feedback controlled, variable speed drive system having an independent regenerating circuitry is shown on Fig.11, which can be considered as th ⁇ further development ( or modification ) of Pig.10.
  • variable speed primary motor 92 of th ⁇ motor vehicle is not connected to shaft 72 - the primary but is driving-y- shaft 98 of a variable displacemenf pump 90.
  • the pressure line 5 ⁇ of pump 90 is connected through check valve 40 to the supply power line L2.
  • variable speed primary motor 92, th ⁇ related speed control block 92 included system.
  • the variable speed primary motor 92 is modulated by th ⁇ Vsupply line pressure drop feedback signal ? 2 — p , which is measured between line 5 (line 91) and line S12.
  • the pump 58 shown on Fig.10 is replaced on Fig.11 by an assisting variable displacement pump 55 having an assisting variable displacement means 57 to make up an assisting supply line pressure drop feedbac N system.
  • the line 36 of pump 55 is connected to tank 62.
  • Th ⁇ pressure line 30 of pump 5 is connected through check valve 44 to line I»2.
  • the assisting ,. variable displacement means 57 i ⁇ modulated by an assisting supply line pressure drop feedback signal P ⁇ — ? Q2 , which is mea ⁇ re ⁇ - line 30 ( through line 32 ) and line SL2.
  • the supply power line L2 is switched over to line 5 or line 30. whichever has th ⁇ higher pressure rate, by th ⁇ logic of ch ⁇ cfc valves 40 and 44. assisting
  • TR ⁇ Vpr ⁇ ssure drop command signal ⁇ 2 ⁇ is selected to be Just slightly larger than en ⁇ Vpr ⁇ sure drop command signal ⁇ ?2 ' Accordingly, while the speed of flywheel is still relatively high, mttiMb* P 2R ⁇ P 02 ⁇ - ⁇ P 2H will . ⁇ ffi*$ $ffir.
  • th ⁇ supply power line L2 will be connected to line 30 through check valve 44.
  • th ⁇ supply power line 2 is connected to line 5- through check valve 40.
  • th ⁇ independent regenerating circuitry including motor 66 t pump 5, and fly-wheel 94, is piven a priority in supplying the fluid energy to the supply power line L2.
  • This independent regenerating circuitry is automatically entering into, and is automatically withdrawing from th ⁇ regulation of fh ⁇ Vsuppi fluid pressure drop across spool valve 2 * Th ⁇ exchange of kinetio energy between the motor vehicle (load 6) and th ⁇ fly-wheel 4 i « basically accomplished as considered above (for th ⁇ schematic shown on Pig.10) t however, th ⁇ undesirable interference between th ⁇ primary motor 92, such as th ⁇ electrical motor or th ⁇ internal-combustion engine, and the regenerating circuitry is now eliminated.
  • variable delivery means 93 of pump 90 can be employed for achieving some additional control objectives, such as maximizing th ⁇ energy efficiency of the internal-combustion engine 92.
  • Fig.11 is of a very general nature and can be further modified and (or) simplified. If there is no additional control objectives, such as just indicated, the variable speed primary motor 92 is replaced by a relatively constant variable for maintaining This case is illustrated by Fig.12 which is a modification of Fig.11 for the hydraulic press type application. In this case, the rotational hydraulic motor 1 is replaced by the double-acting cylinder 1.
  • the motor load and the motor load means are the structural components of any energy regeneratingT ⁇ acTaptive fluid motor control system.
  • Fig.12 (as well as Fig.9 ) also shows the frame 190 ( of a hydraulic press 192 ) ⁇ against which the chamber 10 of cylinder 1 is loaded.
  • the compressed ⁇ bluid energy is basically stored within chamber 10 of cylinder 1; however, the stretching of frame 190 of press 192 may substantially contribute to the calculations of the over-all press energy accumulated under the load.
  • word “LOAD" within block 96 is also considered to be a substitute for the words "the motor load means” and is ⁇ to all ihe possible applications of this invention.
  • the motor load means include a mass of a ".wheeled" motor vehicle ⁇ as it is specifically indicated on the schematic of F ⁇ g.22 ,
  • the fluid motor and load means include the fluid motor means an ⁇ Vmo ⁇ r load means and accumulate a load related energy, such as a kinetic energy of a load mass or a compressed fluid energy of the fluid understood as motor-cylinder.
  • the "exhaust fluid energy” isVa measure of the load related energy being transmitted through the exhaust power line (that is line L3 or line L5) •
  • the “exhaust fluid energy” can also be referred to as the "waste fluid energy',' that is the energy which would be wasted unless regenerated.
  • counterpressure varying means There are basically two types of counterpressure varying means: a) the counterpressure varying means which are not equipped for recupturing the load related energy (such as the exhaust line pressure drop regulator — see figures 1, 4, and 6 ) , a nd b) the counterpressure varying means which are equipped for recupturing the load related energy (such as the exhaust line variable displacement motor — see figures Q, 10, 11, and 12).
  • This counterpressure varying and energy recupturing means can also be referred to as the exhaust line energy recupturing means.
  • the load adaptive fluid motor position feedback control system control/' is typically a combination of at least three component feedback ⁇ systems - the fluid motor position feedback control system, at ieast one exhaust line pressure drop feedback control system, and at least one supply line pressure drop feedback control system.
  • the pressure drop feedback control systems In order to prevent a possible ⁇ mh ⁇ bontiial C fn -CQ, ⁇ . ⁇ . interference between the combj ned components systems, the pressure drop feedback control systems must be properly regulated both with respect to the fluid motor position feedback control system and with respect to each other.
  • pressure-compensated flow characteristics which are shown on figures .3-D, 5-B, and 7-B, can generally be reduced to each of two asymptotic characteristics as follows _
  • a simultaneous speed control of fluid motor 1 by the supply and pressure drop feedback control systems may create avpressure drop regulation interference between these two systems.
  • This pressure drop regulation interference may reveal itself in generating excessive pressure waves, producing hydraulic shocks, cavitating th ⁇ hydraulic fluid, and accumulating an air in the hydraulic tracts.
  • the pressure drop regulation interference may lead IO the over-all dynamic instability of the load adaptive fluid motor control system, such as the regenerative adaptive fluid motor control system.
  • the magnitude and direction oi the motor load is conveniently measured by the pressure signals P Q - ⁇ nnd P 05 - which are implemented for controiling the supply •:ind exhaust line pressure drop leedback control systems, respectively.
  • the load pressure signals J-'n d,ld p 05 are aiso used for controlling the sequence of operation of these pressure drop feedback control systems, as it is illustrated below.
  • the wheeled vehicle is moving with a constant speed.
  • the motor load is positive
  • the load pressure signal P Q 2 is relatively large
  • the pressure signal P 0 5 is very small, and therefore, the exhaust line pressure drop feedback control system is not activated to maintain the exliaust fluid pressure drop P Q g — P. r-r s across spool vaive 2.
  • the exhaust fluid pressure drop P 5 — Pr is equal approximately to the primary supply line pressure drop command signal Z_ o • P rov l clecl that supply and exhaust openings of valve 2 are identical. Note also that
  • the motor load is negative
  • the load pressure signal P 0 5 is large
  • the exhaust line pressure drop feedback control system is activated to maitain the exhaust fluid pressure drop P Q5 — across spool valve 2.
  • the pressure P Q 2 is very small and has a tendency of dropping "below zero".
  • a vacuum in motor line LI must be prevented by introducing a check valve (such as check valve 155 on figures 20 and 22) connecting line LI with the oil tank 62 ( or with a low-pressure hydraulic accumulator).
  • a check valve such as check valve 155 on figures 20 and 22
  • the process of deceleration should be started onle after this check valve is open. It is understood that in this setuation, the supply line pressure drop feedback control systems have no effect on the process of deceleration of motor 1.
  • the motor load is positive
  • the load pressure signal P Q 2 is large
  • the functions of the motor load are not limited to controlling separately each of the pressure drop feedback control systems. Indeed, the functions of the motor load are generally extended to include also the control over the sequence of operation of the supply and exhaust line pressure drop feedback control systems , in order to prevent a possible pressure drop regulation interference between these pressure drop feedback control systems.
  • the load adaptive fluid motor position feedback control system is typically a combination of at least three component feedback control systems - the fluid motor position feedback control system, at least one exhaust line pressure drop feedback control system, and at least o e supply line pressure drop feedback control system.
  • each of the separate component systems is linearized and, thereby, is basically described by the ordinary linear differential equations with constant coefficients, as it is usually done in the engineering calculations of electrohydraulic, hydromechanical, and hydraulic closed-loop systems.
  • the fluid motor position feedback control system (separated from other component systems) is especially easy to linearized if to admit that th ⁇ expected regulation of the exhaust and supply fluid pressure drops is already "in place”.
  • the load adaptive fluid motor position feedback ⁇ ontrol system incorporating only three component systems — the fluid motor position feedback control system, only one exhaust line pressure drop feedback control system, and only one supply line pressure drop feedback control- system.
  • the criterion of dynamic stability of combined component systems can be reduced to only five conditions as follows : (1) providing a dynamic stability of the fluid motcr position feedback control system ;
  • first, second, and third, conditions of dynamic stability are the requirements to the separate component systems.
  • the fourth and fifth conditions of dynamic stability define limitations which must be imposed on the separate component systems in order to combine them together.
  • the design of the separate closed-loop systems for the dynamic stability and required performance is well known in the art, as already emphasized above. For this reason, it is further assumed, for simplicity, that the first three conditions of dynamic stability are always satisfied if the last two- conditions of dynamic stability are satisfied.
  • the last two conditions of dynamic stability are similar, they can also be specified by a general form as follows : preventing a substantial dynamic operation interference between the pressure drop regulation (the exhaust or supply fluid pressure drop regulation) and the motor position regulation by providing a significant dynamic performance superiority for the pressure drop feedback control system (the exhaust or supply line pressure drop feedback control system, respectively ) against the motor position feedback control system.
  • a substantial dynamic operation interference is associated with the concept of providing "a significant dynamic performance superiority".
  • a substantial dynamic operation interference is introduced to characterize the dynamic instability of combined component systems which are stable while separated. This dynamic instability can be detected in a frequency domain or in a time domain by ., M
  • Th ⁇ closed-loop resonant frequency & n ( h& is ⁇ R or C rf ) a located by a resonant peak of th ⁇ closed-loop frequency-response characteristic and, therefore, is also often called "a peaking frequency".
  • This resonant peak is typically observed on a plot of the amplitude portion of the closed-loop frequency-response characteristic. However, the resonant peak is observed. only if the system is underdQ ped.
  • ⁇ oep and ⁇ C d "" the open-loop cross-over frequencies for th ⁇ position feedback control system and th ⁇ pressure drop feedback control system respectively.
  • the final transient time f (that is ⁇ . or t fd ) of » e closed-loop system is tht total output-response time to the step input.
  • the open-loop cross-over frequency Ooc ' and the closed-loop final transient time - are well known in the art — see, for example, the above named books of S. M. Shinners , S. A. Davis , and A.F. D'Souza .
  • T is the minimum stability margin in a frequency domain
  • S ⁇ is the minimum stability margin in a time domain.
  • the final transient ⁇ dimensionless time ⁇ is a function of th ⁇ damping coefficient ⁇ T . More generally, when th ⁇ right part of the second—order equation is more complicated, th ⁇ final transient dimensionless time "S " ⁇ is also effected by th ⁇ right part of this equation.
  • the ratio & ⁇ -Rd can be approximated by the rati.o & ⁇ d and therefore t
  • CO 2 ⁇ ⁇ and ⁇ . are the undamped natural frequency and the f inal transient dimensionless time , respectively , for the position feedback control system j
  • the ratio is basically dependent on the ratio • and is further dependent on th ⁇ secondary factors, such as th ⁇ effects of damping.
  • This main test is needed to prevent th ⁇ frequency resonance type phenomenon between the component systems.
  • the load adaptive fluid position servomechanisms make it possible to substantially improve the energy, performance, and environmental characteristics of the position feedback control in comparison with the conventional fluid position servomechanisms.
  • the load adaptive fluid position servomechanisms may combine the high energy-efficient and quiet operation with the relatively high speed and accuracy of performance.
  • the artificial load adaptability of load adaptive fluid position servomechanisms is achieved by regulating the exhaust and supply fluid pressure drops by the exhaust and supply line pressure drop feedback control systems, respectively.
  • the load adaptive fluid position servomechanisms combine th ⁇ very best qualities of th ⁇ conventional fluid motor position feedback control systems and th ⁇ naturally load adaptive, electric motor position feedback control systems. Moreover, the load adaptive fluid position servomechanisms may incorporate the energy regenerating circuitry.
  • maintaining the exhaust and supply fluid pressure drops across th ⁇ directional control valve may protect th ⁇ position closed-loop against such destructive conditions as generating excessive pressure waves, producing hydraulic shocks, cavitating th ⁇ hydraulic fluid, and accumulating an air in the hydraulic tracts.
  • th ⁇ transition to th ⁇ adaptive servomechanisms makes it easy to control the fluid conditions in the hydraulic tracts and to provide a ' "full hermetization" of th ⁇ hydraulic motor.
  • the load adaptive fluid motor position feedback control systems can be used in machine tools (including presses), construction machinery, agricultural machinery, robots, land motor vehicles, ships, aircrafts, and so on.
  • the load adaptive fluid position servomechanism can be viewed as a combination of a primary motor, such as the electrical motor or the combustion engine, and the load adaptive, position feedback controlled fluid power transmission, transmitting the mechanical power from a shaft of the primary motor to the load.
  • a primary motor such as the electrical motor or the combustion engine
  • the load adaptive, position feedback controlled fluid power transmission transmitting the mechanical power from a shaft of the primary motor to the load.
  • the load adaptive hydraulic press may have advantages against the conventional hydraulic and mechanical presses due to a combination of factors as follows: i. Th ⁇ Veriergy-effici ⁇ ncy of the hydraulic system combining the load adaptive primary power supply and the load adaptive regeneration of energy.
  • the press is easy to control with respect to the moving slide position, stroke, speed, and acceleration.
  • the press maximum tonage is also easy to restrict for the die-tool protection.
  • Fig.13 shows a generalized model of the load adaptive fluid motor output feedback control systems which include an independent energy regenerating circuitry.
  • This model can be viewed as a further development of Fig.8 in view of figures 11 and 12 and is mostly self-explanatory.
  • the position feedoack control means ( block H- ) and t h e related signals X ⁇ , X Q , and X which are shown on Fig.8,* are replaced by the ( motor ) output feedback control means ( block -M ) and the related signals M- j _ , M Q , and ⁇ M, which pre shown on F ⁇ g.l3>
  • the motor position X, , the position input-command signal X , and the position feedback control error signal ⁇ X are replaced by their "generic equivalents” — the motor output M, , the related input-command signal M Q , and tho motor output feedback control error signal ⁇ M, respectively.
  • the motor output feedback control error signal ⁇ M is produced by the output feedback control means (block '- ) in accordance with a difference between the input-command signal H and the motor output M, .
  • the motor output is a generic name at least for the motor position, the motor velocity, and the motor acceleration.
  • the load adaptive fluid motor output feedback control system is a generic name at least for the following systems . a) the load adaptive fluid motor position feedback control system. b) the load adaptive fluid motor velocity feedback control system? c) the load adaptive fluid motor acceleration feedback control system.
  • the general criterion of dynamic stability of combined component systems which was formulated above with respect to the load adaptive fluid motor position feedback control system , is also applicable to the load adaptive fluid motor output feedback control system.
  • the concept of providing "a significant dynamic performance superiority" which ⁇ ⁇ ⁇ -formulated above with respect to the load adaptive fluid motor position feedback control system, is also applicable to the load adaptive fluid motor output feedback control system.
  • the motor velocity V- is the velocity of the moving part 21 of the fluid motor 1.
  • the motor velocity V. can also be viewed as a mechanical signal — the output velocity signal of the load adaptive fluid motor velocity feedback control system.
  • the motor velocity V is measured by the velocity sensor, which is included into block 4-V and is connected to the moving part 21 of the fluid motor 1.
  • the spool of valve 2 is not generally in the neutral spool position but is in the position which corresponds to the given value of the velocity command signal V .
  • the velocity feedback control means (block *4 ⁇ V ) can be still further described basically by the analogy with the above brief description of the position feedback control means (block )•
  • the schematic shown on Fig.16 can be used for constructing the load adaptive , velocity feedback controlled, fluid power drive systems for the motor vehicles.
  • This schematic is derived from the one shown on Fig.11 by replacing the position feedback control means (bloc* ) and the related signals X , X, , and ⁇ X by the velocity feedback control- means (block --V) and the related ri ⁇ nals V , V, , and ⁇ V, respectively.
  • the five-way spool valve 2 shown on Fig.11 is replaced by the four-way spool valve 2 shown on Fig.16.
  • the supply power line L ⁇ and the exhaust power line 3 are eliminated.
  • the four-way spool valve 2 is considered now to be a one- spool directional valve— it's ⁇ aTTbe moved only down from the neutral spool position and can be returned back to the neutral spool position only (which is shown on Fig.l ⁇ ).
  • Regenerative adaptive fluid motor control Regenerative adaptive fluid motor control.
  • FIG.15 A generalized model of the regenerative adaptive fluid motor open-loop control systems is presented by Fig.15 which is derived from
  • Fig.13 just by eliminating the output feedback control means (block - ) and the rel.ated signals M , M, , and ⁇ .
  • the schematics for the load adaptive fluid motor open-loop control systems can be derived from the above presented schematics for th ⁇ load adaptive fluid motor position feedback control systems just by eliminating th ⁇ position feedback control means ( block ) and th ⁇ related signal X Q , X ⁇ , and ⁇ X.
  • the open-loop schematic which is shown on Fig.17 , is derived from the one shown on Fig.16 just by eliminating the velocity feedback control means ( block -V ) and the related signals V , V, , and ⁇ V.
  • Fig.17 The schematic of Fig.17 can be used for constructing the high energy-e ficient load adaptive motor vehicles, as it will be still further discussed later.
  • the general criterion of dynamic stability of combined component systems which was formulated above with respect to the load adaptive fluid motor position feedback control systems, is also applicable to the load adaptive fluid motor open-loop control systems.
  • the concept of providing "a significant dynamic performance superiority" which is formulated above with respect to the load adaptive fluid motor position feedback control system, is also applicable to the load adaptive fluid motor open-loop control system.
  • ⁇ significant dynamic performance superiority of any pressure drop feedback control system against the fluid motor open-loop control system can be established, for example, by providing basically a significantly larger closed-loop bandwidth for this pressure drop feedback control system in comparison with an open -loop cross-over frequency of the fluid motor open- loop control system-.
  • a regenerative adaptive fluid motor control system is typically a combination of at least three component control systems - a fluid motor control system, at least one exhaust line pressure drop feedback control system, and at least one supply line pressure drop feedback control system.
  • the fluid motor control system may or may not include the output feedback control means.
  • the load adaptive motor vehicle drive system is, indeed, an operative regenerative system having all the components working in unison.
  • Adaptive fluid control and the motor vehicles Adaptive fluid control and the motor vehicles.
  • the load adaptiveVvehicle drive systems like the one shown on Fig.17, mny have advantages against the conventional motor vehicle drive systems in terms of such critical characteristics as energy efficiency, environmental efficiency, reliability, controlability, and dynamic performance.
  • the primary supply fluid pressure drop regulation by the variable speed primary motor (engine) 92 has an effect of the energy supply regulation in accordance with the actual energy requirements. , ⁇ , regulation
  • the exhaust fluid pressure drop ⁇ and th ⁇ independent regenerating circuitry make it possible to create th ⁇ schematic conditions, under which th ⁇ energy accumulated during th ⁇ deceleration of th ⁇ motor vehicle is reused during th ⁇ following acceleration of th ⁇ motor vehicle. The energy accumulated during the vehicle down-hill motion will also be reused. .
  • this smaller engine can be substituted by two still smaller engines, only one of which is operated all the time, while the second engine is switched-in only when needed - for example, when the vehicle is moving up-hill with a high speed, as it will be explained more specifically later. 7.
  • the air pollution effect of the motor vehicles will be substantially reduced just by eliminating the waste of energy engines, and brakes-
  • the load adaptive drive system is especially effective in application to the buses which operate within the cities, where a stop-and-go traffic creates the untolerable waste of energy, as well as the untolerable level of air pollution.
  • the bus is moving in a horizontal direction only.
  • the process of bus deceleration - - acceleration beginning from the moment hen the bus is moving with some average constant speed and the "red light" is ahead. Up to this moment the spool of valve 2 have been hold pushed partially down by the driver so that this valve is partially open.
  • the load adaptive drive systems can also be characterized by saying that these drive systems incorporate the energy regenerating brakes.
  • the regenerative adaptive fluid control schematic which is shown on Fig.18, can also be used for the motor vehicle applications, and in particular, for the buses which operate within the cities. This schematic will be studied by comparison with the one shown on Fig.17.
  • the fly—wheel 9 ⁇ shown on Fig.17 is substituted by a hydraulic accumulator 122 shown on Fig.18. Accordingly, the exhaust line variable displacement motor 66 is replaced by the exhaust line constant displacement motor 116 driving tne exhaust line variable displacement pump 120 which is powering the hydraulic accumulator 122 through check valve 136. oU
  • a counterpressure transformer including fluid motor 116, shaft 110, fluid pump 120 , tank lines 74 ' and 134, and power lines 78 and 132, is implemented to make up the counterpressure varying and energy recupturing ⁇ means of the exhaust line pressure drop feedback control system maintaining counterpressure P.
  • the assisting variable displacement pump 55 is replaced by the assisting constant displacement pump 114 being driven by the assisting variable displacement motor 118 which is powered by the hydraulic accumulator 122.
  • Adaptive fluid control the combined energy accumulating means.
  • Fig.19 which is basically a repitition of Fig.18, however, two major components are added, the electrohydraulic energy converting means 142 and the electrical accumulator 144.
  • variable speed primary motor 92 is replaced by the constant speed primary motor 100, so that now the variable displacement mechanism 93 of pump 90 is used for regulating the supply fluid pressure drop P ? — , as it was already illustrated by Fig.12.
  • the hydraulic accumulator 122 is almost fully charged, an excess fluid is released from this accumulator, and a hydraulic energy of the excess fluid is converted through the electrohydraulic energy converting means 142 to the electrical energy of electrical accumulator 144.
  • the schematic of Fig.19 can be characterized by that the combined energy accumulating (and storing) means include the fluid energy accumulating means being implemented for powering the electrical energy accumulating means. More generally, the combined energy accumulating (and storing) means may include major (primary) energy accumulating means being implemented for powering supplementary (secondary) energy accumulating means.
  • the combined energy accumulating (and storing) means may include fluid energy accumulating means (hydraulic accumulator 122 on Fig.19 ) being implemented for powering the electrical power line (replacing electrical accumulator 144 on Fig.19 ).
  • the electrical power line will accept an excess energy from the hydraulic accumulator 122 and will return the energy back to the hydraulic accumulator 122 — when it is needed.
  • Fig.20 is basically a repetition of Fig.10; however, trie variable speed primary motor 9 is introduced now by the variable speed primary Internul-combussion engine 92.
  • driving load 96 is replaced by a variable displacement motor 15 n driving the same load.
  • variable displacement means 152 of motor 150 are constructed to make-up the displacement feedback control system including a variable displacement mechanism (of motor 150 ) and employing a displacement feedback control errow signal AD, is generated in accordance with a difference between a spool displacement ( command signal ) D of valve 2 and a mechanism displacement ( feedback signal ) D- ⁇ of the variable displacement mechanism of motor 150.
  • the displacement feedback control system which is well known in the art, is, in fact, the position feedback control system and that, therefore, the general position feedback control technique, which is characterised above with respect to the fluid motor position feedback control system, is also basically applicable to the displacement feedback control 'system.
  • the higher speed range is produced by changing the displacement of motor 150.
  • the lower speed range of motor 150 is defined between the "zero" spool position and the point of full actual ( orifice ) opening of valve 2. Up to this point, the command signal D is kept constant, so that the displacement of motor l ⁇ O is maximum and is not changed.
  • the higher speed range of motor 150 is located beyond the point of full actual ( orifice ) opening of valve 2. Beyond this point ( due to the spool shape of valve 2 ) the further spool displacements do not change any more the opening of valve 2. On the other hand, beyond this point, the oomm n ⁇ alg-r d. D is being reduced b,y tne i'ur ⁇ ner spool displacements of valve Z. Accordingly, tne displacement ___ j ⁇ — D of tne variable displacement mechanism of 1 o motor 150 is being also reduced by tne displacement feedback control system. The smaller the displacement of motor 150, the higher the speed of this motor ( and tne smaller the available torgue of this motor ) .
  • Fig. 0 also illustrates the use of check valves for restricting the maximum and minimum pressures in the hydraulic power lines.
  • the check valve 154 is added to very efficiently restrict the maximum pressure in the exhaust motor line L4 by relieving an excess fluid from this line ( through check valve 15 ⁇ ) into the high-pressure hydraulic accumulator 122.
  • the check valve 155 is added to effectively restrict the minimum pressure in the supply motor line Ll by connecting this line ( through check valve 155 ) with the tank 62. ote that tank 62 can generally be replaced by a low-pressure hydraulic accumulator ( accompanied by a small— supple etary tank ) .
  • the available braking torque should be usually substantially larger than the available accelerating torque.
  • Fig.21 is basically a repetition of Fig.18 ; however, the constant displacement motor 1 driving the load 96 is also driving a regenerative braking variable displacement pump 17 0 which is used to increase the available regenerative braking torque.
  • the tank line 176 of pump 170 is connected to tank 62.
  • the pressure line 178 of pump 170 is connected through check valve 17 to the hydraulic accumulator 122.
  • the flow output of pump 170 is regulated in accordance with the pressure rate P ⁇ in the motor line L4 conducting a m o tor fluid flow from the fluid motor 1, as it is more specifically explained below.
  • ⁇ pressure-displacement transducer converting the pressure signal
  • Proportional command-displacement signal d — C *P nt - is included into the variable displacement means 99 of pump 170.
  • This transducer may incorporate, for example, a small spring-loaded hydraulic cylinder actuated by the pressure signal p °5*
  • displacement feedbabk control system which is well known in the art, is, in fact, the position feedback control system and that, therefore, the general position feedback control technique, which is characterised above with respect to the fluid motor position feedback control system, is also basically applicable to the displacement feedback control system.
  • the displacement d.. of pump 170 is increasing acdordingly, so that the total regenerative braking torque is properly distributed between the fluid motor 1 and the regenerative braking pump 170.
  • the variable displacement motor 110 ynd the constant displacement pump 114 are replaced by the constant displacement motor 190 and the variable displacement pump 19>I, i n order to provide a wider r a nge of regu l ation of pressure.
  • the assisting constant displacement motor 198 Is powered by the hydraulic accumulator 122 (through shut-off valve 299 ) and is driving the assisting variable displacement pump 194 which is pumping the oil from tank 62 back into the accumulator 122 (through check valve 204 and shut-off valve 299 ) .
  • the output flow rate of accumulator 122 (in line 210) is equal to a difference between the input flow rate of motor 198 (in line 200) and the output flow rate' of pump 194 (in line 1 4).
  • Time torque of pump 194 counterbalances the torque of motor 190.
  • the pressure P ?R in line 30 can be regulated from "almost 7,ero" ⁇ ,o the •• maximum” , accordingly.
  • the check valve 208 connects Line 132 ( of pump 120 ) with the tank 62.
  • the shut-off valve 29J? is controlled by the load pressure signal P Q2 •
  • the check valve 208 and shut-off valve 2 9 9 are considered to be optional and are introduced only to illustrate more specifically some exemplified' patterns of controlling the load adaptive exchange of energy between the fluid mo Lor and load means and the energy accumulating means. The related explanations are presented below .
  • This initial pump displacement is made just slightly negative, in order to provide for the pump 120 a very small initial output ( in line 13 ⁇ ) directed to the tank 62, and thereby, to provide for the exhaust fluid flow (in line L5 ) a free passage through motor 116 to the tank 62.
  • the check valve 208 is open, the cneck valve 136 is closed, 3tnd the pump 120 is actually disconnected from the accumulator 122.
  • the motor vehicle is moving in a horizontal direction, up-hill, and down-hill, and with the different speeds, accelerations, and decelerations; however, all what counts for controlling the energy recupturing pressure drop feedback control system, is the load rate and direction (which are measured by the pressure signals P n ,- and P 02 )• While the pressure signal P & c . is very small, the pump 120 is actually disconnected from the accumulator 122, and the exnaust fluid flow is passing freely through motor 116 to the tank 62. As tne pressure signal P ⁇ e is increasing, t h e kinetic energy of a vehicle mass is converted to the a ccumulated energy of accumulator 122.
  • the first major modification is identified by using the variable speed primary motor 92 for regulating the primary supply fluid pressure d rop, as illustrated by figures 11, 16, 17, 18, 20, and 21.
  • the second major modification is identified by using the variable displacement mechanism of the variable displacement primary pump 90 for regulating the primary supply fluid pressure drop, as illustrated by figures 12, 19, and 22. It is important to stress that these two major modification ⁇ are often convertible.
  • the schematics shown on figures 11, 16 , 17, 18, 20, and 21 can be modified by replacing the v a riable speed primary motor 92 by a constant speed prim a ry motor 100 and by using the variable displacement primary mechanism of pump 90 for regulating the 'supply fluid pressure drop P 2 — * as it is illustrated by figures 12, 19, and 22.
  • transition to the modified schematics is further simplified by providing a constant speed control system for the variable speed motor 92 and by converting, thereby, this variable speed motor to a constant speed motor.
  • the motor vehicle is first accelerated by actuating tne variable displacement mechanism of pump 90 — as illustrated ana by Fig.227V"Ts further accelerated by actuating the variable speed primary interrial-oombussion engine — as illustrated by
  • the motor vehicle is first accelerated .by actuating tne variable displacement mechanism of pump 90 — as illustrated by Fig. 2, is further accelerated by actuating the variable speed primary i nternai-oombus ion engine — as illustrated by Fig. 0, and is still further accelerated by actuating the variable displacement mechanism of motor 15 — as illustrated by figures 20 ⁇ iid 22.
  • the engine will be usually fully loaded only during the third stage of speed regulation — just after the displacement of motor 1 is sufficiently reduced.
  • the minimum possible displacement of motor 150 must be restricted by the desirable maximum of engine load (which can be measured, for example, by the desirable maximum of pressure P ft? in line Ll of motor 150 ).
  • the motor vehicle is first accelerated by actuating the variable displacement mechanism of pump 90 — as illustrated by Fig.22, and is further accelerated by actuating the variable speed primary internal-combussion engine — as illustrated by Fig. 0. Contrary to point 2, there is no third consecutive stage of speed regulation ( by using the variable displacement motor 150 ). Instead, the displacement of motor 150 is controlled independently by using the pressure signal P Q2 which is provided by line Ll. The larger the pressure signal P n? , the larger the displacement of motor 150 — within the given limits, of course, 4.
  • the motor vehicle is provided with two relatively small engines. The first engine time. The second engine is while the motor vehicle io moving up-hill with a high speed. Enc engine is driving * sep rate pump ( like pump 90 ). Each engine-pump instalation is working with a separate spool valve ( likn ⁇ pool vnlvo 2 ).
  • the first option of operation ( rsee point ] ) is npplied to the ⁇ econd engine-pump n ⁇ t nl t ion of the two-engine
  • the independent regenerating circuitry can be easily switched-off by the driver in the process of operating a motor vehicle. This can be accomplished by using a directional valve switching over the exhaust power line L5 from the energy regenerating circuitry to the tank.
  • regenerative adaptive drive system such as shown on Fig.22, can be modified by replacing the "stationary" exhaust line energy recupturing means (the constant displacement motor 116 driving the variable displacement pump 120) and the "stationary" assisting variable delivery fluid power supply
  • the energy regenerating, load adaptive drive system of a wheeled vehicle can be still further modified to provide an optional mechanical connection of the engine shaft with the wheels of the vehicle.
  • This optional mechanical connection can be used, for example, for long-distance driving.
  • the de*sign of modified-integrated drive system may include an integrating mechanical transmission to select one of two alternative - component systems as follows:
  • the basic regenerative adaptive drive system see figures 17 to 22.
  • the engine of a vehicle is connected with the primary pump 90.
  • the back axil of ⁇ vehicle is driven by the constant displacement motor 1 (or by the variable displacement motor 150).
  • the optional conventional power train In this case, the shaft of the engine is connected mechanically to the back axil of a vehicle.
  • Regenerative adaptive fluid motor control the energy recuperating pressure drop feedback control system.
  • a regenerative adaptive fluid motor control system having an independent regenerating circuitry (see figures 11 to 22 ) p ⁇ is an integrat*ng- system incorporating only two major components i * Q n) the two-way load adaptive fluid motor control system which is adaptive to the motor load along the exhaust and supply power lines of the- spool valve 2, and b) the two-way load adaptive energy regenerating system which is also adaptive to the motor load along the exliaust and supply power lines of the spool valve 2.
  • the regenerative system having an independent regenerating circuitry is charactirized by that the primary and assisting supply line pressure drop feedback control systems are separated.
  • the exhaust line pressure drop feedback control system (which can prevail also be referred to d the energy recupturing pressure drop reedback control system ) is shared between the two-way load adaptive fluid motor control system a nd the two-way load adaptive energy regenerating system.
  • the energy recupturing pressure drop feedback control system includes an exhaust line energy recupturing means for varying a counterpressure rate in the exhaust power line and for recupturing a load related energy, such as a kinetic energy of a load mass or a compressed fluid energy of a fluid motor-cylinder.
  • the energy recupturing pressure drop feedback control system and the exhaust line energy recupturing means can also be referred to as the energy recuperating pressure drop feedback control system and the exhaust line energy recuperating means, respectively.
  • the fluid motor and load means include the fluid motor means and the motor load means and accumulate a load related energy (such as a kinetic energy of a load mass or a compressed fluid energy of the fluid motor-cylinder ) for storing and subsequent regeneration of this load related energy.
  • a load related energy such as a kinetic energy of a load mass or a compressed fluid energy of the fluid motor-cylinder
  • exhaust fluid energy of the exhaust fluid flow is understood as a measure of the load related energy being transmitted through the exhaust power line (that is line L3 or line L5).
  • the “exhaust fluid energy” can also be referred to as a “waste fluid energy”, that is the energy which would be wasted unless regenerated .
  • the f rs e'nergy converting means include the energy recupturing pressure drop feedback control system and convert the load related energy of the fluid motor and load means to an accumulated energy of the energy accumulating means for storing and subsequent use of this accumulated energy.
  • the high energy—efficient, load adaptive process of converting the load related energy to the accumulated energy is facilitated by regulating the exhaust fluid pressure drop across spool valve 2 by the energy recupturing pressure drop feedback control system and is basically controlled by the motor load. Note that the energy is being accumulated by the energy accumulating means, while the motor load is negative
  • the secon ⁇ Venergy converting means include the assisting supply line pressure drop feedback control system and convert the accumulated energy of the energy accumulating means to an assisting pressurized fluid stream being implemented for powering the supply power line L2 of spool valve 2.
  • the assisting- pressurized fluid stream is actually gene ' rated by an assisting variable delivery fluid power supply which is included into the assisting supply line pressure drop feedback control system and which is powered by the energy accumulating means.
  • the high energy-efficient, load adaptive process of converting the nccu ulated energy to the assisting pressurized fluid stream is facilitated by regulating the assisting supply fluid pressure drop across spool valve 2 by the assisting supply line pressure drop feedback control system and is basically controlled by the motor load. Note that the energy is being released by the energy accumulating menns , while the motor load is positive ( for example, during the acceleration of the motor vehicle ).
  • the load adaptive energy regenerating system is also basically controlled by the motor load.
  • the primary supply line pressure drop feedback control system includes a primary variable delivery fluid power supply generating a primary pressurized fluid stream being implemented for powering the supply power line L2 of the spool valve 2.
  • the assisting supply line pressure drop feedback control system includes an assisting variable delivery fluid power supply generating an assisting pressurized fluid stream being also implemented for powering the supply power line L2 of the sppo.l valve 2.
  • the assisting pressurized t huid stream has a priority over the primary pressurized rluid stream in supplying the fluid power to the supply power line L2.
  • regeneration of a load related energy of the fluid rn ⁇ tor and load means is accomadated by correlating the primary pressure rate of the primary pressurized fluid stream with the assisting pressure rate of the assisting pressurized fluid stream by regulating the primary supply fluid pressure drop across valve 2 and regulating the assisting supply fluid pressure drop across valve 2 by the primary supply line pressure drop feedback control system and the assisting supply line pressure drop feedback control system, respectively.
  • the exhaust line energy recupturing means of the energy recupturing pressure drop feedbyck control systems can be introduced by the exhaust line variable displacement motor
  • the assisting variable delivery fluid power supply which is powered by the energy accumulating means, can be introduced by the ' assisting variable displacement pump 55 — see figures , 11, 12, 16, 1.7, or by. the assisting variable displacement motor 118 driving the assisting constant displacement pump 114 — see figures 18 to 21.
  • the assisting variablle delivery fluid power supply can also be introduced by the assisting constant displacement motor 198 driving the assisting variable displacement pump 194 - as it is illustrated by Fig.22.
  • the primary variable delivery fluid power supply can be introduced by the primary variable displacement pump 90 - see figures 12, 19, and 22 or by the variable speed primary motor
  • any pressure drop regulation is accomplished by the related pressure drop feedback control system by implementing the related pressure drop feedback signal for modulating one of the following : a) the variable displacement means of the variable displacement pump , b) the variable displacement means of the variable displacement motor, c) the variable speed primary motor ( or the variable speed primary engine ) driving the primary fluid pump.
  • variable displacement pumps having the built-in pressure drop feedback controllers are well known in the art. This type of control for the 'variable displacement pump is often called a "load sensing control" and is described in many patents and publica tioiros ( see, for example, Budzich — —U.S. Patent No , 07*., 529 of Feb.21, I978 ).
  • variable displacement pumps wit h t h e load-sensing pressure drop feedback controllers are produced (in mass amount ) by many companies which provide catalogs and other information on this load sensing control. Some of these companies are : a) THE OILGEAR COMPANY, 2300 South, 51th Street, Milwaukee, W I 3219, U.S.A. ( see, for example, Bulletin 47016A ) ; b ) SAUER-SUNDSTRAND COMPANY, 2800 East 13th Street, ⁇ mes I ⁇ 50010, U.S.A. ( see, for example, Bulletin 9825, Rev.E ) ,- c) DYNEX/ IVETT , INC., 770 Capitol Drive, Pewaukee,
  • the load adaptive variable displacement means ( of the variable displacement pumps and the variable displacement motors ), which are used " in this invention, are basically similar with the well-known load-sensing pressure drop feedback controllers of the variable displacement pumps. These load adaptive displacement means can also be reffered to as the load adaptive displacement controllers.
  • Re ic h swas h plate is ⁇ driven " by a ⁇ unge ⁇ bTl:. e related cylinder against the force of a precompressed spring.
  • Each hydraulic cylin d er is controlled by the related three-way spool valve which is also provided wj ⁇ th the pressure and tank lines.
  • the pressure line is powered by an input pressure P which is supplied by any appropriate pressure sourse.
  • the valve spool is driven by a pressure drop feedback signal against the force of the precompressed spring defining the pressure drop command signal.
  • three-way valve can also be replaced by a two-way val e which does not have the tank line ( in " this case the tank line i connected through a throutle ⁇ al vs to the. line of hydraulic cylinder) .
  • the spool 252 of valve 250 is in the neutral spool position which is shown on Fig.23. Note that • ⁇ . was already indicated before.
  • F ig.25 illustrates an examplified energy regenerating circle. It is assumed that the wheeled vehicle is moving in a horizontal d irection only. As the vehicle is moving with a constant speed , decelerated, completely stoped, and accelerated, the related energy regenerating circle is completed. This stop-and-go energy regenerating circle has been already briefly introduced before ( to explain the concept of preventing a substantial pressure drop regulation interferrence ) and is easily readable on Fig. 5, when considered in conduction with figures 22 to 24 and the related text. For example, while the vehicle is decelerated, the swashplate 266 is positioned as indicated on Fig.24. While the vehicle is accelerated, the swashplate 246 is positioned as indicated on Fig.23.
  • Regenerative drive system having the combined energy accumulating means.
  • the schematic shown on Fig..19 is now further modified to replace the independent regenerating circuitry by the built-in regenerating circuitry and to improve the utilization of the combined energy accumulating means. Accordingly, the assisting variable delivery fluid power supply (motor 118 driving pump 114), the check valves 40 and 44, and the electrohydraulic energy converting means 142 are eliminated.
  • the modified schematic is shown on Fig.26. The added components are:
  • the regenerative drive system of Fig.26 makes it possible to minimize the required engine size of a wheeled vehicle.
  • the engine 100 is provided with a speed control system which is assumed to be included in block 100 and which is used to maintain a preselected (basic) speed of shaft 98 while allowing some speed fluctuations under the load which is applied to the shaft 98.
  • the related margin of accuracy of the speed control system is actually used to maitain a balance of power on the common shaft 98 and, thereby, to minimize the required engine size of a wheeled vehicle.
  • the driving torque of shaft 98 is generally produced by engine 100, by motor-generator 290 (when it is working as a motor), and by motor 300 (when it is powered by the hydraulic accumulator 122 through shut-off valve 298).
  • the loading torque of shaft 98 is basically provided by pump 90 and by motor-generator 290 (when it is working as a generator). Note that at some matching speed of shaft 98 (within the margin of accuracy of the speed control system) a speed-dependent voltage of generator 290 is equal to a charge-dependent voltage of accumulator 144, so that no energy is transmitted via lines 292 and 294. As the speed of shaft 98 is slightly reduced, the electrical energy is transmitted from the electrical accumulator 144 to the electrical motor 290 helping engine 100 to overcome the load.
  • shut-off valve 298 is normally closed and is open only under some preconditions - in order to power the constant displacement motor 300 by the hydraulic energy of accumulator 122.
  • the pump 90 is basically powered by engine 100.
  • the pump 90 is basically powered by motor 300 and is also powered by engine 100 and motor 290.
  • the constant displacement motor 300 is powered by the hydraulic accumulator- 122, through shut-off valve 298.
  • ⁇ n optional control signal "S" which is applied to the shut-off valve 298, is produced by an optional control unit which is not shown on Fig. 6.
  • This control unit can be used for controlling such optional functions as follows:
  • variable displacement pump 90 is also used as a motor to provide an al ternative route for transmission of energy from accumulator 122 to the common shaft 98;
  • regenerative adaptive fluid motor control systems There are basically two types of regenerative adaptive fluid motor control systems: (a) the regenerative system having an independent regenerating circuitry (see figures 11 to 22) and (b) the regenerative system having a built-in energy regenerating circuitry (see figures 9, 10. and 26).
  • the first type of regenerative systems is identified by that the primary and assisting supply line pressure drop feedback control systems are separated.
  • the second type of regenerative systems is identified by that the primary and assisting supply line pressure drop feedback control systems are not separated and are represented by only one supply line pressure drop feedback control system.
  • the generalized first-type systems have been already introduced by figures 13, 14, and 15.
  • ⁇ generalized second- ype system is shown on Fig.27, which is mostly self-explanatory and is still further understood when compalred with figures 9, 10, 26, and 15. Note that transition from the first to the second type of regenerative systems is accomplished typically by replacing the separated primary and assisting supply line pressure drop feedback control systems by only one supply line pressure drop feedback control system and by implementing the primary power supply means for powering the energy accumulating means.
  • the transition from the independent regenerating circuitry to the built-in regenerating circuitry can be accomplished by eliminating the separated primary supply line pressure drop feedback control system and by implementing the primary pump 90 for powering the hydraulic accumulator 122 (the resulted schematic can be still further modified to incorporate also an electrical accumulator ) .
  • the two basic types of regenerative systems can generally be combined to include both - the built-in regenerating circuitry and the independent regenerating circuitry.
  • the transition to the combined schematic can be accomplished by adding an assisting supply line pressure drop feedback control system, which is shown on Fig.22 and which includes the constant displacement motor 198 driving the variable displacement pump 194.
  • the resulted combined schematic is also applicable to the wheeled vehicles .
  • Th ⁇ above .load-related classification of typical adaptive schematics is instrumental in modifying these schematics for tho modified load environments.
  • the schematic shown. on Fig.18 is adaptive tc t h e two-directional dynamic load force, which is generated during acceleration and deceleration of a load mass moving only in one direction.
  • the schematic of Fig.15 must be also modified.
  • the modified schematic may include he five-way spool valve 2 instead of the four-way spool valve 2 which is shown on Fig.18.
  • the energy regenerating circuitry using hydraulic accumulator 122 must be switched over from the exhaust o e line L5 to the exhaust power line L .
  • the regenerative braking pump 170 of Fig.21 can also be used as a variable displacement motor to make-up a supplementary variable displacement motor/pump.
  • the pump functions of this supplementary output motor/pump have been already studied with the help of Fig.21.
  • the motor functions of this supplementary output motor/pump will also be studied - separately.
  • Fig.28 is derived from Fig.21 by replacing the supplementary output pump 170 by the supplementary output motor 170 and by eliminating the assisting supply line pressure drop feedback control system (including motor 114 and pump 118 ) and some other components (check valves 40, 44, and 174).
  • the variable displacement motor 170 is powered by the hydraulic accumulator 122 through a shut-off valve 297 which is basically controlled • by pressure signal Pg 2 . While this pressure signal is comparatively small, the shut-off valve 297 is closed. ⁇ s signal P Q 2 I S further raising-up, the shut-off valve 297 is open, provided that there is still enough energy stored in the hydraulic accumulator 122.
  • variable displacement means 99 of motor 170 are constructed to make-up a displacement feedback ⁇ control system including a variable displacement mechanism (of motor 170) a displacement feedback control errow signal ⁇ d, generated in accordance with a difference between
  • command-displacement signal do C » P n 0 ⁇ 2 (where Cp is a constant coefficient ) and a mechanism displacement ( feedback signal ) d. of the variable displacement mechanism of motor 170, ⁇ pressure-displacement transducer converting the pressure signal P Q 2 into the proportional command-displacement signal d Q? is included into the variable displacement means 99 of motor 170.
  • Tli transducer may incorporate, for example, a small spring-loaded hydraulic cylinder actuated by the pressure signal p
  • the displacement feedback control system which is well known- in the art, is, in fact, the position feedback control system and that, therefore, the general position feedback control technique, which is characterised above with respect to the fluid motor position feedback control system, is also basically applicable to the displacement feedback control system.
  • motor line Ll is further ralsing-up, the displacement d of motor 170 is increasing acdordingly, so that the total accelerating torque is properly distributed between the fluid motor 1 and the supplementary motor 170.
  • the use of motor 170 makes it possible to substantially increase the available (total) accelerating torque of the wheeled vehicle.
  • Fig.29 The related generalized schematic of Fig.29 is derived from Fig.15. is mostly self-explanatory, and is reflective of the facts that the assisting supply line pressure drop feedback control system is now eliminated and that pressure P ac from the hydraulic accumulator 122 is now applied to the supplementary output motor 170 of the fluid motor and load means.
  • Fig.26 The schematic of Fig.26 can be modified by changing the hybrid motor means driving pump 90.
  • the examplified modifications are as follows.
  • the electrical motor-generator 290 and the related electrical accumulator 144 are exluded from this schematic.
  • the constant displacement motor 300 is replaced by a variable displacement motor which is used to construct a supplementary shaft-speed feedback control system maintaining the preinstalled speed of shaft 98 when this variable displacement motor is powered by accumulator 122.
  • the hydraulic energy of accumulator 122 is transmitted to shaft 98 in accordance with the actual energy requiren ent . Note that possible interference between the main shaft-speed feedback control system ( of primary engine 100 ) and the supplementary shaft-speed feedback control system ( of the variable displacement motor ) is prevented by providing
  • V CM — is a velocity command-signal for the main shaft-speed feedback control system
  • V cs is a velocity command—signal for the supplementary shaft-speed feedback control system
  • ⁇ V — is a sufficient velocity margin between these two systems.
  • the supplementary speed control system should actually be regulated just "slightly above” the main speed control system.
  • the primary engine 100 is excluded from the schematic of Fig.26. In this case, the primary energy should be supplied by the electrical accumulator 144.
  • the primary engine 100 is disconnected from shaft 98 and is driving a constant displacement pump which is powering the constant displacement motor 300.
  • the hydraulic energy of accumulator 122 is transmitted to shaft 98 via this constant displacement pump driving the constant displacement motor 300.
  • the schematic of Fig.22 can be modified by providing the primary engine 100 with a variable-speed feedback control system which is used for maintaining the engine maximum energy efficiency. Note that as the engine speed increases, the displacement of pump 90 is being reduced accordingly, to maintain the pump flow output which is defined only by the opening of valve 2.
  • Fig.22 can also be modified by eliminating the primary supply line pressure drop feedback control system
  • the resulted schematic having a built-in energy regenerating circuitry can also be constructed for maintaining the engine maximum energy efficiency.
  • the primary supply power line 54 (see figures .1.1 to 22) can be protected by the maximum pressure relief valve.
  • the maximum pressure in line 54 can also be restricted by using the variable delivery means 93 of pump 90.
  • the maximum pressure relief valves can also be used to protect other hydraulic lines.
  • the check valve 154 (figures 20 and 22) is added to very efficiently restrict the maximum pressure in the exhaust motor line 1,4 by relieving an excess fluid from this line (through check valve 154) into the high-pressure hydraulic accumulator 122.
  • the check valves can be used to restrict the maximum pressure in still other power lines.
  • the check valve 155 (figures 20 and 22) is added to effectively restrict the minimum pressure in the supply motor line hi by connecting this line (.through check valve 155) with the tank 62. ' .
  • check valves can be used to restric the minimum pressure in still other power lines.
  • the exhaust power line 1.5 (or L3) should usually be connected through a check valve to the tank to avoid creating a vacuum in this line.
  • the oil tank capacity can often be reduced, the oil cooling system can often be eliminated.
  • the oil tank 62 can often be replaced by a low-pressure hydraulic accumulator (accompanied by a small-supplementary tank ) .
  • the oil tank 62 can also be supplemented by a low-pressure centrifugal pump.
  • the practical design must include the means of restricting the spool displacement (SD) of valve 2 versus the load pressure rate (LP) in line Ll ( or in line L2 ), so that the resulted load power rate (which is proportional to LP x SD ) would not exceed the limited power supply capacity.
  • the practical regenerative adaptive fluid motor control systems may include the means of restricting the required load power rate in accordance with the limited power supply capacity.
  • the non-regenerative adaptive fluid motor control systems are equipped with an exhaust line pressure drop feedback control system including an exhaust line pressure drop regulator.
  • the regenerative adaptive fluid motor control systems are equipped with an energy recuperating pressure drop feedback control system including an exhaust line energy recuperating means .
  • the schematic shown on Fig can be easily modified to convert the five-way valve 2 to the six-way valve by separating the supply power line L6 from the supply power line L2. The separated line L6 can be then connected directly to the line 5 ⁇ of the additional hydraulic power supply 5° shown on Fig. .

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  • Physics & Mathematics (AREA)
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Abstract

A regenerative adaptive fluid motor control system integrating a load adaptive fluid motor control system and a load adaptive energy regenerating system having an energy accumulator. The fluid motor control system includes a primary variable displacement pump (90) powering a spool valve (2) controlling a fluid motor (1) accumulating a load related energy. The load related energy of the fluid motor is regenerated to provide a load adaptive exchange of energy between the fluid motor and the energy accumulator (122). This load adaptive exchange of energy is combined with a load adaptive primary energy supply for maximizing the over-all energy efficiency and performance potentials of the fluid motor control. The load adaptability is achieved by regulating the exhaust and supply fluid pressure drops across the spool valve.

Description

(54) REGENERATIVE ADAPTIVE FLUID CONTROL
(75) Invenl;σrι Robei-t M. Lisniangky , Broolclyn, Mew York.
(73) Assignee i
(21) International Appln No.
(22) Filed!
(63) This is related to U.S. application S.N.08/715, 434 of of 09/18/96 ( to be issued as a patent )
WHICH IS A CIP OF 08/399, 123 03/06/95 ABAN
WHICH IS A CIP OF 08/075, 288 I 36/11/93 ABAN
WHICH IS Λ CIP OF ' Ό7./.VΛ ICH IS Λ ciP. .or 07/521 i ,6w63" i"i73ϊ/9. ΛDΛN
WH 05/10/90 ΛBΛN
WHICH- IS Λ CIP or 07/301 ,616 01/25/89 ΛBΛN
WHICH IS Λ CI or 07/096 , I20 09/11/07 ΛI3ΛM
WHICH IS Λ DIV or 06/737 ,063 05/23/05 ΛUΛM
WHICH IS Λ CI or 06/701 ,325 02/ I 3/05 ΛDΛN
WHICH IS Λ CIP or 06/310 ,672 I 1/05/0I ΛI3ΛN
( 51 ) Infc.Cl1
(52) u.s.ci 60/325, 60/327, 60/328, 60/388,
60/393, 60/413, 6o/4l4, 6o/428, 6o/445, 60/459, 60/494, 60/686, 60/706, 100/35, 100/48; 100/46,
(58) Field of Search...60/325% 327, 328, 388, 393, 413, 414, 428, 445, 459, 494, 686, 706, 100/35, 48, 46, 180/165, 308, 60/419, 468X, 414XR, 494XR, 437XR, 417/214, 292XR, 91Λ00, 5l4XR, 451Xι 520XR, 137/106XR, 596.1XR
(56) References Cibed
U.S. PATENT DOCUMENTS 4,118,149 10/1978 Hagberg ... 060/431 3.882,896 5/1975 Budzich 09.1/446
4,364,229 12/1982 Shiber 060/414
4,693,000 9/1987 Van Hooff 060/417
2,924,940 2/1960 Covert et al 060/430
OTHER PUBLICATIONS
Lisniansky, Robert M., "Avtomatika e Regulirovanie Gidravlicheskikh pressovV Moskow: Mashinostroenie, 1975.
REGENERATIVE ADAPTIVE LUID CONTROL
This is related to U.S. application Serial No.08/715, 434, filed 09/18/96, to be issued as a patent, which is a continuation-in-part of application Serial No. 08/399,123, filed 03/06/95, now abandoned. which is a continuation-in-part of application Serial No, 08/075,288, filed 06/11/93, now abandoned. which is . a continuation-in-part of application Serial No. 07/815,175, filed 12/31/1991, now abandoned, which is a continuation-in-part of application Serial No. 07/521,663, filed 5/10/1990, now abandoned, which is a continuation-in-part of application Serial No. 07/301,646, filed 1/25/1989, now abandoned, which is a continuation-in-part of application Serial No. 07/096,120, filed 9/14/1987, now abandoned, which is a divisional of application Serial No. 06/737,063, filed 5/23/1985, now abandoned, which is a continuation-in-part of application Serial No. 06/704,325, filed 2/13/1985, now abandoned, which is a continuation-in-part of application Serial No. 06/318,672, filed 11/05/81, now abandoned. FIELD OF THE INVENTION
The present invention relates primarily to a fluid motor position feedback control system, such as the electrohydraulic or hydromechanical position feedback control system, which includes a fluid motor, a primary variable displacement pump, and a spool type directional control valve being interposed between the motor and the pump and being modulated by a motor position feedback signal. More generally, this invention relates to the respective fluid motor output feedback control systems and to the respective fluid motor open-loop control systems. In a way of possible applications, this invention relates, in particular, to the hydraulic presses and the motor vehicles. My copending international application on Regenerative Adaptive Fluid Control is identified by Serial No.
The larger picture of the Energy-Regenerating Adaptive Fluid Control Technology is ..presented by my six U.S. applications identified by Seεi$l Numbers' 08/715,470; 08/716,474;" 08/715,434; 08/7lO"i-3'23; 08/710,567; 08/725,056.
BACKGROUND ART: TWO MAJOR PROBLEMS
The hydrauJ lc Cluid motor is usually driving a variable load. In the variable load environments, the exhaust and supply fluid pressure drops across the directional control valve are changed, which destroys the linearity of a static speed characteristic describing the fluid motor speed versus the valve spool displacement. Al a rβault, a system fain and tht related qualities, such as tht dynamic performance and accuracy, are all tht function* of tht variable load. Moreover, an energy efficiency of tht petition feedback control is also a function of the variable load.
The more the load rate and fluctuations, and the higher the performance requirements, the more obvious are the limitations of the conventional fluid motor position feedback control systems. In fact, the heavy loaded hydraulic motor is especially difficult to deal with when several critical performance factors, such as the high speed, accuracy, and energy efficiency, as well as quiet operation, must be combined. A hydraulic press is an impressive example of the heavy loaded hydraulic motor-mechanism. Tht load conditions are changed substantially within each press circle, including approaching the work, compressing tht fluid, tht working stroke , dβcompretaing tht fluid, and tht rtturn stroke.
A more comprehensive study of the conventional fluid motor position feedback control systems can be found in numerous prior art patents and publications — see, for example : a) Johnson, J.E. , "Electrohydraulic Servo Systems", Second Edition. Cleveland, Ohio . Penton/jPC , 1977. b) MerritVH.E. , "Hydraulic Control. Systems" .
New York - London - Sydney : John Wiley & Sons, Inc., 1967. c) Lisniansky, R.M., "Avto atika e Regυlirovanie Gidravlicheskikh Presgov".
Moscow : Mashinostroenle, .1975
(this book had been published in Russian only).
The underlying structural weakness of the conventional fluid motor position feedback control systems can be best characterized by saying that these systems are not adaptive to the changing lsa environments.
The problem of load adaptability of the conventional βiβctrohydraulic and hydromechanical position feedback control systems can be more specifically identified by analyzing t*o typical hydraulic schematics.
The first typical hydraulic schematic includes a three- ay directional control valve In combination with the two counteractive (βxpansiblβj chambers. The first of these chambers is controlled by said three-way valve «hich is alto oonntcttd to the pressure and tank lines of the fluid power means. The second chamber Is under a relatively constant pressure provided by said pressure lint. In this cast, It Is not possible to automatically maintain a supply fluid pressure drcp across the three-way valve without a "schematic operation interference- *ith the position feedback control system. Indeed, maintaining the supply fluid pressure drop can be achieved only by changing tht prtssurt lint pressure, which is also applied to tht second chamber and, therefore, must be kept approximately constant.
The second typical schematic includes a four-way directional control valve in combination with the two pountβractive chambers. Both of thtst chambers are controlled by the four— way valve which is also connected to the pressure and tank lints of tht fluid power means. In this schematic, it is not possible to automatically maintain an exhaust fluid across the four-way valve without encountering which can also' be viewed as a schematic operation interference with the position feedback control system. Indeed, a chamber's pressure signal which Is needed for maintaining the exhaust fluid pressure drop, must be switched over from one chamber to the other in exact accordance with a valve spool transition through a neutral spool position, where the chamber lines are switched over, to avoid damaging the spool valve flow characteristics. In addition, a pressure differential between the o chambers at the neutral spool position *U1 affect the pressure drop regulation and may generate the dynamic unstability o the position feedback ccntrol system.
■ still The problem of load adaptability can ST' further Identified by emphasizing a possible dynamic operation inτerferer. e bβt.vββn the position feedback control and the regulation of the exhaust and supply fluid pressure drops.
The problem of load adaptability can be still further identified by emphasizing a possible pressure drop regulation interference between the supply and exhaust line pressure drop feedback control systems.
The structural weakness of the conventional fluid motor position feedback control systems can be still further characterized by that these systems are not equiped for regenerating a load related energy, such as a kinetic energy of a load mass or a compressed fluid energy of the fluid motor-cylinder. As a result, this load related energy is normally lost. The problem of load adaptive regeneration of energy is actually correlated with the problem of load adaptability of the fluid motor position feedback control system, as it will be illustrated later.
Speaking in general, the problem of load adaptability and the problem of load adaptive regeneration of energy are two major and interconnected problems which are to be solved consecutively by this invention, in order to create a regenerative adaptive fluid motor position feedback control system and, finally, in order to create a regenerative adaptive fluid motor output feedback control system and a regenerative adaptive fluid motor open-loop control system.
SUMMARY OF THE INVENTION performance The present invention is primarily aimed to improve the ^ qualities and energy efficiency of the fluid motor position feedback control system, such as the elβctrohydraulic or hydromechanical position feedback control system, operating usually in the variable load environments.
The improvement of performance qualities, such as the dynamic performance and accuracy, is the first concern of this invention, while the improvement of energy efficiency is the second but closely related concern.
This principal object is achieved by »
(a) shaping and typically linearizing the flow characteristics of the directional control valve by regulating the supply and exhaust fluid pressure drops across this valve ;
(b) regulating the hydraulic fluid power delivered to the directional control valve, in accordance with,but above, what is required by the fluid motor i
(c) preventing a schematic operation interference between the regulation of said pressure drops and the position feedback control >
(d) preventing a dynamic operation interference bβt.vββn the regulation of said pressure drops and the position feedback control (as it will be explained later);
(e) preventing a pressure drop regulation interference between the supply and exhaust line pressure drop feedback control systems (as it will also be explained later).
The implementation of these interrelated steps and conditions is a way of transition from the conventional fluid motor position feedback control systems to the load adaptive fluid motor position feedback control systems. These load adaptive systems can generally be classified by the amount of controlled and loadable chambers of the fluid motor, by the spool valve design configurations, and by the actual shape of the spool valve flow characteristics. In a case hen only one of two counteractive chambers cf the fluid motor is controllable, the fluid motor can be loaded only n one direction. The controlled chamber Is connected to the three-way spool valve which also has a supply pow«r line and an exhaust power line. In this case, the second chamber is under a relatively constant pressure su plied by an independent source of fluid power.
In a case when both chambers are controllable, the fluid motor can be loaded in only one or in both directions. The controlled chambers are connected to a fivβ-*.ay spool valve which also has a common supply power line and two separate exhaust power lines. *'hen the fluid motor is loaded in only one direction, 'only one of two exhaust lines is aJLso a cbuntβrprβssure line. n'hen the fluid motor is loaded in both directions, both exhauat llnββ are used as oountβrprβssurβ lines.
Using the thrββ-w*y or five-way spool valve with a separate exhaust line for each controllable chamber, makes it possible to prevent a schematic operation interference between the position feedback control and- the regulation of pressure drops. In particular, the problem of measuring a chamber's pressure signal is eliminated. Each countβrpressurβ line is provided *ith an exhaust line pressure drop regulator. l*hich is modulated by an exhaust line pressure drop feedback signal which is measured between this counterpressure line and the r»l.atji,d_ chamber.
In tht process of maintaining tht supply fluid pressure drop across tht spool valvt, a supply fluid flo* rate is being monitored continuously by the primary variale displacement pump of the fluid power means. Maintaining the supply fluid pressure drop is also a way of regulating the hydraulic power delivered to the spool type directional control valve. In the process of maintaining the exhaust fluid pressure i rz p across the spool valve, all the flow is being released from the counterpressure line through the exhaust line pressure drop regulator to the tank. Counterpressure may be created in the counterpressure line only for a short time while the hydraulic fluid in the preloaded chamber is being decompressed. However, .the control over the decompression is critically Important for improving the system's dynamic performance potential.
A family of load adaptive fluid position servomechanisms may include the three-, four-, five-, and six-way directional valves
The three-way spool valve is used to provide the individual pressure and counterpressure lines for only one controllable chamber. The six-way spool valve is used to provide the separate supply and exhaust lines for each of two controllable chambers.
The five-way spool valve can be derived from the six-way spool valve by connecting together two separate supply lines.
The four-way spool valve can be derived from the five-*ay spool valve by connecting together -two separate exhaust lines.
The four-way spool valve does create a problem of schematic operation interference between the position feedback control and the regulation of pressure drops, as it is already explained above.
However, the principal possibility of using the four-way spool valve in the adaptive position servomechanisms Is not excluded.
What, is in common for the adaptive fluid position servomechanisms being considered is that the fluid motor is provided with at least one controlled and loadable chamber, and that this chamber is provided with the pressure-compensated spool valve flow characteristics. These pressure-compensated flow characteristics are shaped by the related exhaust line pressure drop feedback control system which includes the exhaust line pressure drop regulator and by the related supply line pressure drop feedback control system which Includes the primary variable displacement pump.
The deqired (linear or unlinear) shape of the spool valve flow characteristics is actually implemented by programming the supply and exhaust line pressure drop command signals of the supply and exhaust line pressure drop feedback control systems, respectively. Some possible principals of programming these command signals are illustrated below.
(1) The supply and exhaust line pressure drop command signals are set approximately constant for linearizing the pressure-compensated spool valve flow characteristics. The related adaptive hydraulic (electrohydraulic or hydromechanical) position servo echanismsc can be referred to as the linear adaptive servomechanisms, or as the fully-compensated adaptive servomechanisms. Still other method of programming the piessure drop command signals can be specified with respect to the linear adaptive servomechanisms, as it is illustrated below- -by points 2 to 5.
(2) The supply line pressure drop command signal is being increased slightly as the respective load pressure rate is increased, so that to provide at least some over-compensation aJong the supply power line.
(3) The supply line pressure drop command signal is being reduced slightly as the respective load pressure rate is increased, so that to provide at least some under-co pensat ion along the supply power line.
(4) The exhaust line pressure drop command signal is being increased slightly as the respective load pressure rate is increased, so that to provide at least some under-compensation along the exhaust power line.
(5) The exhaust line pressure drop command signal is being reduced slightly as the respective load pressure rate is increased, so that to provide at least some over-compensation along the exhaust power line.
It is understood that the choice of flow characteristics do not effect the basic structure and operation of the load adaptive fluid motor control systems. For this reason and without the loss of generality, in the following detailed description, the linear adaptive servomechanisms are basically considered.
It is a further object of this invention to develop a concept of load adaptive" regeneration of a load related energy, such as a kinetic energy of a load mass or a compressed fluid energy of the fluid motor-cylinder. This is achieved by replacing the exhaust line pressure drop regulator by a counterpressure varying and energy recupturing means ('such as an exhaust line variable displacement motor or an exhaust line constant displacement motor driving an exhaust line variable displacement pumpK by replacing the exhaust line pressure drop feedback control system by an energy recupturing pressure drop feedback control system, and finally, by creating a load adaptive energy regenerating system including fluid motor and load means and energy accumulating means.
I is still further object of this inventiion to develop a concept of load adaptive exchange of energy between the fluid motor and load means and the energy accumulating means of the load adaptive energy regenerating system. The load adaptive regeneration of the load related energy of the fluid motor and load means can be viewed as a part (or as a larger part) of a complete circle of the load adaptive exchange of energy between the fluid motor and load means and the energy accumulating means.
It is still further object of this invention to develop a regenerative" adaptive fluid motor position feedback control system which is an integrated system combining the load adaptive fluid motor position feedback control system and the load adaptive energy regenerating system.
It is still further object of this invention to develop a regenerative adaptive fluid motor output feedback control system and a regenerative adaptive fluid motor open-loop control system. In general, the regenerative adaptive fluid control makes it possible to combine the load adaptive primary power supply and the load adaptive regeneration of energy for maximizing the over-all energy efficiency and performance potentials of the fluid motor control systems.
1 is still further object of, this invention to develop the high energy-ef icient, load adaptive hydraulic presses utilising the regenerative adaptive fluid control.
It is still further object of this invention to develop the high energy-efficient, load adaptive motor vehicles utilizing the regenerative adaptive fluid control.
It is still further object of this invention to develop the high energy-efficient, load adaptive City Transit Buses utilizing the regenerative adaptive fluid control.
Further objects, advantages, and futures of this Invention will be apparent from the following detailed description when read in conjuctlon with the drawings.
BRIEF DESCRIPTION OF THE DRAWINGS Fig.l sho , thβ adaptive fiuid 3βrvomecha*i3m having only one controllable chamber. Pig.2 shows a power supply schematic version. Fig-3-Λ is a generalization of Fig.l.
Fig. -D illustrates the flow characteristics of valve 2. Piβ. shows the adaptive fluid sβrvomβchanlsm having two controllable chambers but loadable only in one direction, Fig.5-Λ is a generalization of Fig. .
Fig.5-B Illustrates the flow characteristics of valve 2. Fig.6 shows thβ adaptive fluid sβrvomechanism having two controllable chambers and loadable in both directions. Flg.7-Λ is a generalization of Fig.6.
Flg.7-n illustrates the flow characteristics of valve Z . rig.θ shows a generalized model of adaptive fluid position servomechanisms . Fig. illustrates the concept of load adaptive regeneration of energy. Fig.10 shows the adaptive fluid servomechanism having
. a built-in energy regenerating circuitry. Fig.11 shows the adaptive fluid servomechanism having an independent energy regenerating circuitry.
Flp..l2 is a modification of Fig.H for the hydraulic press type applications. Fig.13 shows a generalized model of the regenerative adaptive fluid motor output feedback control systems. Fig.l'. shows a generalized model of the regenerative adaptive fluid motor velocity feedback control systems. Fig.15 shows a generalized model of the regenerative adaptive fluid motor open-loop control systems. Fig.iό is a modification of Fig.11 for the .motor vfehicle type applications. Fig.17 shows a regenerative adaptive drive system for the motor vehicle type applications. Fig. IB shows a regenerative adaptive drive system having a hydraulic accumulator. Fig. 9 shows a regenerative adaptive drive system having the combined energy , regenerative means. Fig.20 shows a regenerative adaptive drive system having a variable displacement motor driving the load. Irig.2) shows a regenerative adaptive drive system having a regenerative braking pump. Fig.22 shows a modified regenerative system having a hydraulic accumulator. Fig.23 shows the load adaptive displacement means of the assisting supply line pressure drop feedback control system.
Fig.24 shows the load adaptive displacement means of the energy recupturing pressure drop feedback control system. Fig.25 illustrates a stop-arid-go energy regenerating circle. Fig.26 shows a modified regenerative system having the combined energy regenerating means. Fig.27 shows a generalized regenerative system having a built-in regenerating circuitry. Fig.28 shows a regenerative adaptive drive system having a supplementary output motor. Fig.29 shows a generalized regenerative system having a supplementary output motor. DESCRIPTION OF THE INVENTION
GENERAL LAYOUT AND THEORY
Introduction: Adaptive fluid position feedback control. Fig.l shows a simplified schematic of the load adaptive fluid motor position feedback control system having only one controllable chamber. The moving part 21 of the fluid motor-cylinder 1 is driven by two counteractive expansible chambers - chambers 10 and 11, only one of which - chamber ID - is controllable and can be loaded. The second chamber - chamber 11 - is under a relatively low (and constant) pressure PQ supplied by an independent pressure source. This schematic is developed primarily for the hydraulic press type applications. Λs it is already mentioned above, the load conditions are changed substantially within each press circle including approaching the work, compressing the fluid (in chamber 10), the working stro k, decompressing the fluid (in chamber 10), and the return strock.
The schematic of Fig.l further includes the hydraulic power supply means 3-1 having a primary variable displacement pump powering the pressure line 51., The three-way spool-type directional control valve 2 is provided with three hydraulic power lines Including a motor line — line LI - connected to line 15 of chaβbβr 10 , thβ supply power line L2 connected to pressure line 51 . and the exhaust power line LJ. Lines L2 and IQ are commutatβd *lth line LI by thβ spool valve 2. To consider all the picture, Fig.l should be studied together with the relate -supplementary figures Z , 3-Λ, and 3-13.
Thβ block represents a generalized model of the optional position feedback control means. This block is needed to actually make-up thβ fluid motor position feedback control .system, which is capable of regulating the motor position X, of motor 1 by employing the motor position feedback signal CX^, where coefficient "C" is, usually, constant. Thβ motor position feedback signal CX, is generated -by a motor position sensor, which is included into block '»■ and is connected to the moving part 21 of thβ hydraulic fluid motor 1.
An original position feedback control error signal ΔXQr is produced aa a difference between the position input-command signal X and the motor position feedback signal CX^. There are at least two typical fluid motor position feedback control systems — the electrohydraulic and hydromechanical position feedback control systems. In the electrohydraulic system, thβ equation —CX,l o the like is simulated by electrical means located within block .
In the hydromechanical system, the equation X„ o — CX.1 or the like is simulated by mechanical means located within block - .
The block may also include the electrical and hydraulic amplifiers, an electrical torque motor, the stabilization— — optimization technique and other components to properly amplify and condition said signal ΔXQr for modulating said valve 2. In other words, the original position feedback control error signal Δx is finally translated • into a manipulated position feedback control error signal ΔX which can be identified with the valve spool displacement Δx from the neutral spool position Δx = 0. manipulated In general, it can be -said that theVposition feedback control error signal Δx is derived in accordance with a difference between the position input-command signal X and the output position signal X.. At the balance of the position feedback 0 and,, hence, Δχi≥0. plicity, it can also be often assumed that Δ X =-X — CX-, .
This principal characterization of the optional position feedback control means is, in fact, well known in the prior art and will be extended to still further details later. 10
Thβ exhaust line pressure drop regulator 3-3 is introduced to make up the exhaust line pressure drop feedback control system which is capable of regulating the exhaust fluid pressure drop across valve 2 by varying thβ counterpressure rate P., in thβ exhaust po-vβr linβ'L3. This exhaust fluid pressure drop is represented by thβ exhaust line pressure drop feedback signal, *hich is equal ?Q-> — P3 and is measured between thβ exhaust power line L3 and thβ related exhaust signal line SL3 connected to line LI. Thβ regulator 3-3 is connected to thβ exhaust power line L3 and to thβ tank line 52 and is modulated by an exhaust line pressure drop feedback control error signal, which is produced in accordance with a difference between thβ exhaust line pressure drop command signal Δ. P_ and thβ exhaust line pressure drop feedback signal P _ — P .
The primary variable displacement pump of fluid power supply means 3-1 ( pump 59 on Pig.2 ) is introduced to make-up thβ supply line pressure drop feedback control system, which is capable of regulating the supply fluid pressure drop across valve 2 by varying thβ pressure rate P2 in the supply power line L2 by varying thβ supply fluid flow rate in said line L2 by said variable displacement pump. This supply fluid pres¬ sure drop is represented by the supply line pressure drop feedback signal, which is equal P2—' PQ2 *rΔ is measured between line L2 (through line 32 on Pig.2) and thβ related sϋpSiX line 3L2 connected to lint LI. A variable delivery means 56 of pump 58 is modulated by a supply line pressure drop feedback control error signal, which is produced in accordance with a difference between thβ supply line pressure drop command signal ΔP2 *nd the supply line pressure drop feedback • signal ?2 — P02 •
The schematic shown on Pig.l operates as follows. At thβ balance of thβ motor position feedback control 1 Δ == rt-CX, = 0. When thβ hydraulic fluid motor 1 is moving from thβ one position X, to thβ other, thβ motor speed is defined by thβ valve spool displacement ΔXaβX0— CX. from thβ neutral spool position ΔxaO. Thβ system performance potential is substantially improved by providing thβ linearity of thβ spool valve flow characteristic P=r K, Δx , '••Here . is the constant coefficien ,and ? is thβ fluid flow rate to ( P. ) or thβ fluid flow rate from ( P t ) thβ controllable chamber 10. This linearity is achieved by applying thβ supply line pressure drop command signal drop command pressure lir.e pressure drop feedback control system, respectively.
Thβ pressure maintained in thβ supply power line L2 by thβ supply line pressure drop feedback control system is - ΔP2 *nd can be Just slightly above what is require for chamber 10 to overcome the load. On the other hand, the counterpressure maintained in the exhaust power line L3 by the exhaust line pressure drop feedback control system is P03 """"ΔP3 *nd can be Just slightly below the pressure PQ- —r PQ2 in chamber 10 . However, there are some limits for acceptable reduction of the pressure drop command signals ΔP? *nd •
The pressure drop command signals ΔP2 and ΔP • the pressure PQ and their interrelationship are selected for linearising thβ spool valve flow characteristic ( PSBK, ΔX ) without "running a risk" of full decompressing the hydraulic motor ( chamber 10 ) and generating thβ hydraulic shocks in the hydraulic system. Some of the related considerations are 1
1. The pressure PQ has to compress the hydraulic fluid in chamber 10 to such an extent as to prevent the full decompression under the dynamic operation conditions. In the absence of static and dynamic loading, the pressure P10 in chamber 10 is fixed by thβ pressure PQ applied to l o
chamber 11 so that KΛ O PΛ O , Where Krt 0 is the constant coefficient. s selected
Under this condition,thβ pressure drop ?0 — P^ as ΔP-j can be maintained even during the return stroke.
Indeed, after decompressing the preloaded chamber 10, thβ regulator 3-3 is open ( P- = 0 ) , but thβ pressure drop
Prt„ — 0 s=s ΔPT •=• K PΛ is still maintained simply by
03 J 0 0 approximately constant pressure PQ .
3. If thβ passages of thβ spool valve 2 are symmetrical relative to thβ point Δx ~ 0, thβ pressure drop command
where 1 P 2min 9 *** mlnimua pressure rate maintained in line L2 by the supply line pressure drop feedback control system. . The smaller pressure drop command signals P and ΔP^ , the larger spool valve 2 is required to conduct thβ given fluid flow rate.
Thβ regulator 3-3 is opened by a force of thβ spring shown on Fig.l and is being closed to provide thβ counterpressure P-j only after thβ actual pressure drop PQ, — p_ exceeds its prβinstallβd value Δ . which is defined by the spring force Practically, at thβ very beginning of thβ return stroke, *hβn the regulator 3-3 has to enter into thβ operation, thβ controllable chamber 10 is still under thβ prβββurβ. It means that regulator 3-3 is preliminarily closed and is ready to provide thβ countβrprββsure P. , which is being maintained by regulator 3-3 only for a short time of decompressing chamber 10. However, the control over the decompression is critically important for improving thβ system's dynamic performance potential. The schematic of Fig. 2 is a disclosure of block 3-1 shown on Fig.l. This schematic includes the primary variable displacement pump 58, which is connected through line 30 and check valve 44 to the pressure line 51. A relatively low pressure, high capacity fluid power supply 50 ( such as a centrifugal pump) is also connected through line 5- and check valve O to the pressure line ji. Thβ primary motors ( such as electrical motors driving the pumps are not sho- on Pig.2. Thβ variable delivery means 56 of pump 58 includes a variable displacement mechanism of this pump. The tank lines 52 and 36 are collected by the oil tank 62. The pressure line 51 can be protected by the maximum pressure relief valve *hich is not sho*n on Pig.2. Thβ maximum pressure in line 5 can also be restricted by using the variable delivery means 5 of pump 58. In general, the maximum pressure relief valves can also be used to protect other hydraulic lines.
In accordance with Fig.2, a relatively low pressure fluid from thβ high capacity fluid power supply 50 is introduced through check valve ^0 into thβ pressure line 51 to increase thβ speed limit of thβ hydraulic cylinder 1 ( Fig.l ), as thβ pressure rate in line 1 is sufficiently declined. Actually, thβ hydraulic po/.er supply 50 is being entered into thβ operation Just after thβ spool of valve 2 passes its critical point, beyond vhich the pressure P- in line 51 is dropped bβlo* the minimum regulated pressure P2min '
Thβ schematic shown on Fig.l is a sy mβtrical, relative to the chambers 10 and 11. The functional operation of this schematic can be still better visualized by considering its generalized model, hich is presented on ?ig.3-A and is accompanied by thβ related pressure-compensated flow characteristic P-^K-^Δx of valve 2. Thβ fluid power means 3 shown on Fig.3-A, combine the fluid power supply means 3-1 and the regulator.* 3-3 , which are shown on Fig.l- The concept of preventing a substantial schematic operation interference.
Fig. shows a simplified schematic of the load adaptive fluid motor position feedback control system having two controllable chambers but loadable only in one direction. This schematic is also developed primarily for the hydraulic press type applications, is provided with the five-way spool valve 2, and is easily understood when compared with Fig.l. The line 12 of chamber 11 is connected to line L4 ol valve 2. The loadable chamber 10 is controlled as before. The chamber 11 is commutated by valve 2 with the supply power line Lb and with the "unregulated" separate exhaust line L5. The supply power line L6 is connected to line L2 but is also considered to be "unregulated", because the supply signal line SL2 is communicated (connected) only with chamber 10. The exhaust line L5 is, in fact, the tank line. In this case, equation (1) can be generalized as:
Δ>2 as, Δ »3 - p 10= u = -^s— • <2 >
where t P,Q and P,, are the pressures in chambers
10 and 11, respectively, at thβ absence of static and dynamic loading.
Thβ pressures P1Q and P^ have to be hi^h enough to prevent thβ full decompression of chambers 10 and 11 under thβ dynamic operation conditions. On thβ other hand, thβ pressure drop command signals Δ 2 and Δ i have to be small enough to improve thβ system energy efficiency.
Thβ schematic shown on Fig is assymmβtrical, relative to thβ chambers 10 and 11. Thβ functional operation of this schematic can be still better visualized by considering its generalized model, which is presented on Fig.5-A and is accompanied 'by thβ related flow characteristics ^I ""1 κl ^* and flow power means 3 shown on Fig.5-A, combine the fluid power supply means 3-1 and the regulator 3-3 » which are shown on Fig . 4 .
The schematic shown on Fig.6 is related to the load adaptive hydraulic position servomechanism having two controllable chambers and loadable in both directions. This schematic is provided with thβ five-way spool valve and is easily understood when compared with Fig. . The loadable chamber 10 is controlled as before except that the supply signal line SL2 is communicated ( commutated ) with chamber 10 through. check valve 5. Thβ second loadable chamber — chamber 11 — is commutated by valve 2 with thβ supply power line ,L6 and with thβ exhaust power line L5. Thβ line L6 is connected to line L2. Thβ supply signal line SL2 is also communicated ( commutated ) with chamber 11 through check valve 6.
Thβ exhaust line L5 is a separate counterpressure line which is provided with an additional exhaust line pressure drop feedback control system including an additional exhaust line pressure drop regulator 3-k which is shown on Fig.6.. The related exhaust signal line SL5 transmitting signal PQ5 ,1s connected to line 12 of chamber 11. Thβ counterpressure maintained in line L5 by the additional exhaust line pressure drop feedback control system, is ι P* =s PQ5 P< . where ΔP^ is thβ related pressure drop command signal.
Thβ check valve logic makes it possible for thβ line SL2 to select one of two chambers, whichever has thβ higher pressure rate, causing no problem for maitaining the supply fluid pressure drop across valve 2, as well as for the dynamic stability of the fluid motor position feedback control system. A very small throttle valve 19 connecting line SL2 with the tank line 52, is helpfull in extracting signal PQ2 .
Thβ schematic shown on Fig.6 is symmetrical, relative to thβ chambers 10 and 11. The functional operation of this schematic can be still better visualized by considering its generalized model, which is presented on Pig.7-A and is accompanied by thβ related pressure-compensated flow characteristics F10= Kj_ ΔX and Fu= — Kχ ΔX of valve 2. The fluid power means 3 shown on Fig.7 -A,combine thβ fluid power supply means 3-1. the regulators 3-3. 3-1*. and the small throttle valve 19»which are shown on Fig.6.
Of course, the linear flow characteristics shown on Fig.3-B, Fig.5-B, and Fig.7-B, are only the approximations of the practically expected flow characteristics of valve 2 , while they are not saturated.
The motor load which is not shown on the previous schematics, is applied to the moving part 21 of thβ hydraulic fluid motor 1. This load is usually a variable load, In terms of its magnitude and (or) direction, and may generally include the static and dynamic components. Thβ statio loading components are thβ onβ-dirβctlonal or two-directional forces.
The dynamic (inertia) loading component is produced by accelerating and decelerating a load mags ( inciudin the mass of moving part 21) and is usually a two-directional force. If thβ fluid motor 1 is loaded mainly only in one direction by a static force, the schematic of Fig.l or Fig.l* is likely to be selected.
If thβ fluid motor 1 is loaded substantially in both directions by thβ static forces , the schematic of Pic.6 is more likely to be used.
What is in common for schematics shown on Fig.3-Λ, Fig.5-A, a d Fig.7-A, is that fluid motor 1 is provided with at least one controlled and loadable chamber, and that this chamber is provided with the pressure-compensated spool valve flow characteristics. This idea can be best illustrated by a model of Fig.8 which is a generalization of Fig.3-A, Fig.5-A, and Fig.7-A. The block 5 of Fig .8 combines fluid motor means (the fluid motor 1) and spool valve means (the spool valve 2), which are shown on previous schematics.
It is understood that load adaptive fluid motor position feedback control systems being considered are not limited to the hydraulic press type applications. As the supply and exhausty lines L2, L3, L5, L6 are commutated with the chamber lines LI L4, the related signal lines SL2, SL3, SL5, SL6 must be communicated accordingly with thβ same chamber lines 11, L^.
The communication of signal Unβs SL2, SL3, S15, SL6 with thβ chambers can be provided by connecting or commutating these signal lines with thβ chambers. Having thβ separate supply and exhaust power lines for each controllable chamber, as well as having only one loadable chamber, makes it possible to eliminate the need for commutating these signal lines.
Finally, it can be concluded that:
1. Providing a separate exhaust power line for each controllable chamber is a basic precondition for preventing a substantial schematic operation interference between the pressure drop feedback control systems and the fluid motor position feedback control system . This schematic operation interference may lead to the dynamic instability of the fluid motor position feedback control system, as it was already explained before.
2. By virtue of providing the separate exhaust power lines L3 and L5, the need for commutating the related signal line SL3 and SL5 is eliminated, as it is illustrated by figures
4 and 6.
3. In a case of having only one controllable chamber, the commutation of supply signal line SL2 is not needed, as it is illustrated by Fig.l.
4. in a case of having only one loadable chamber, the commutation of supply signal line SL2 can be avoided, as it is illustrated by Fig. .
5. In a case of having two loadable chambers, the commutation of supply signal line SL2 can be accomplished by such commutators as follows:
(a) the commutator using check valves 5 and 6 and being operated by the pressure differential between the power lines of motor 1, as it is illustrated by Fig.6 >
(b) " the commutator using an additional directional control valve which is operated by the spool of valve 2.
6. In accordance with point 5, the schematic of Fig.6 can be modified by replacing the first-named commutator by the second-named commutator.
The modified schematic is of a very general nature and is applicable to the complex load environments.
Position feedback control means.
It should be noted that transition from the conventional fluid position servomechanisms to the load adaptive fluid position servomechanisms does not change the part of the system which is outlined by block . The optional physical structure of the position feedback control means is disclosed in numerous prior art patents and publications describing thβ conventional fluid motor position feedback control systems and the related position feedback control technique — see, for example, the above named books and also » a) Davis, S.A., and B.K. Lβdgerwood, "Electromechanical Components for Servomechanisms? New York : McGraw-Hill, 1961 b) Wilson, D.R., Ed., "Modern Practice in Servo Design? Oxford-New York-Toronto-Sydney-Braunschweig ι
Perga on Press, 1970. c) Analog Devices, Inc., "Analog-Digital Conversion Handbook", Edited by Sheingold D.H., Third Edition. Englewood Cliffs, N.J.i Prentice-Hall, 1986. d) D'Souza, A.F., "Design of control systems". Enplewood Cliffs, N.J.i Prentice-Hall, 1988-
It should also be noted that the electrical position feedback control circuitry of electrohydraulic position servomechanisms is quite similar to that of electromecha- nicaϊvg rvonβcTianisms. It is to say that in the case of electrohydraulic position servomechanisms, the electrical portion of block — including thβ optional position sensor but excluding thβ electrical torque motor — can also be characterize by the analogy with thβ comparable portion of thβ electric motor position feedback control systems — — see, for example, thβ books already named above.
In accordance with thβ prior art patents and publications, thβ above brief description of block is further emphasized and extended by the comments as follows J
1. The motor position X^ is the position of moving part 21 (piston, shaft and so on) of the fluid motor 1. in fact, the motor, position X^ can also be viewed as a mechanical signal — the output position signal of the fluid motor position feedback control system being considered.
2. The motor position Xj is measured by the position feedback control means due to the position sensor, which is included into block ^ and is connected to the moving part 21 of thβ fluid motor 1.
3. In the electrohydraulic position servomechanisms, an electromechanical position sensor can be analog or digital. The analog position sensor , employs an analog transducer, such as a linear variable differential transformer, a synchro transformer, a resolver and so on. The digital position sensor may include a digital transducer, such as an optical encorder. The digital positionsensor can also be introduced by an analog-digital combination, such as the resolver and the resolver-to-digital converter — see, for example, chapter l of thβ above named book of Analog devices, Inc. . It is to say that in the electrohydraulic, analog or digital, position servomechanisms, thβ motor position feedback signal CX, ( or thβ like ) is generated by the electromechanical sensor in a form of thβ electrical, analog or digital, signal, respectively. __,,.
26
5. It is also to say that in thβ electrohydraulic, analog or digital, position servomechanisms, thβ position input - -command signal X is also thβ electrical, analog or digital, signal, respectively. Thβ position input-command signal X can be generated by a variety of components— from a simple potentiometer to a computer.
6. In thβ hydromechanical position servomechanisms, thβ mechanical position sensor is simply a mechanical connection to thβ moving part 21 of thβ fluid motor 1. In this case, thβ motor position feedback signal CX, is a mechanical signal. Thβ, position input-command signal X is also a mechanical signal.
7. in accordance with explanations given previously .
(a) the original position feedback control error signal ΔX is produced as a difference between the position input-command signal X and the motor position feedback signal CX-_ 1
(b) the original position feedback control error signal ΔXor is finally translated into the manipulated position feedback control error signal Δx ;
(c) it can be said that the manipulated position feedback control error signal ΔX is derived in accordance with a difference between the position input-command signal X and the output position signal X, >
(d) the manipulated feedback control error signal ΔX is a mechanical signal, which is identified with the spool displacement of valve 2 from the neutral spool position ΔX=0
8. in the electrohydraulic position servomechanisms, the spool of valve 2 is most often actuated through the hydraulic amplifier of thβ position feedback control means. The spool valve 2, the hydraulic amplifier, and the electrical torque motor are usually integrated into what is called an "electrohydraulic servovalve".
9. In thβ hydromechanical position servomechanisms, the spool of valve 2 is also most often actuated through the hydraulic amplifier of the position feedback control means. Thβ spool valve 2 and thβ hydraulic amplifier are usually integrated into what is called a "sβrvovalvβ? 10. Still more comprehensive descriptioon of the optional position feedback control means ( block ) can be found in the prior art patents and publications including thβ books already named above.
A concept of load adaptive regeneration of energy.
thβ compressed hydraulic fluid is substantial in defining the system energy efficiency, a regeneration of this energy can ba Justified. Pig* is originated by combining Fig.l and Fig.2. However, thβ regulator 3-3 is replaced by a variable displacement actor 6 having a variable displacement means 67 t a pressure line 77, and a tank line 7
Thβ motor 65 is connected through line 77 to line L3 and has a •cαemwn shaft" 72 with thβ variable displacement pump 58.
Thβ variable displacement means 67 is modulated by thβ
(through line 75) and thβ related signal line SL3. feedback Thβ exhaust line pressure drdj contfol system Including motor 65, maintains the exhaust fluid pressure drop TQ -> — P, across spool valve 2 by varying thβ counterpressure '3^ '03 —" ΔF3 in the exhaust line L3 by thβ variable displacement means 67 .
A fly-wheel 9 is attached to the shaft 72 and is driven by motor 65. The pump 58 is generally driven JyTN motor 100, by the motor 65 and by .
As a result, thβ potential energy of tht fluid compressed in chamber 10 and, hence, the exhaust fluid energy of thβ exhaust fluid flow passing through line L3 , is converted into a kinetic energy 0f motor 65 and thβ related rotated mass including fly-wheel 9**. Thia kinetic energy is finally 2o
reused through thβ supply power line L2 by thβ supply line pressure drop feedback control system.
Fig.9 also shows the frame 190 ( of hydraulic press 192 ) against which the chamber 10 of cylinder 1 is loaded.
The concept of load adaptive regeneration of energy is further illustrated by considering the load adaptive, position feedback controlled, variable speed drive systems for the motor vehicle type applications (see figures 10 and 11), where a kinetic energy associated with a mass of the motor vehicle is substantial in defining the over-all energy efficiency. It will be shown that load adaptability of these efficient and flexible drive systemSj makes it easy to create the schematic conditions under which the energy accumulated during decelerating the motor vehicle is reused for accelerating the vehicle.
It is understood that availability of the motor position inpu -command signal X0 makes it possible not only to regulate the fluid motor position X , but also to control the lluid motor velocity. It is now assumed, for simplicity, that motor vehicle is moving only in a horizontal direction. Accodingly, it is also assumed that five-way spool valve 2 is working now as a one-directional valve - it's spool can be moved only down from the neutral spool position and can be returned back to the neutral spool position only (which is shown on figures 10 and 11)' Note that figures 10 and 11 are used only for a further study of
Load adaptive regeneration of energy. The related veloci y feedback control (Fig.16) and especially the related open-loop control (figures 17 to 22, and 26) are, of course, more likely to be used for the motor vehicle type applications.
In general, the load adaptive, position feedback controlled, variable speed drive systems may incorporate a built-in regenerating circuitry or an independent regenerating circuitry. The drive system incorporating the built-in regenerating circuitry is shown on Fig.10 which is originated by combining Fig.6 and Fig.2. However, the fluid power supply of Fig.2 is represented on Fig.10 mainly by pump 58. The regulator 3-3 is not needed now and, therefore, is not shown on Fig.10. On the other hand, the regulator 3-4 is replaced by a variable displacement motor 66 having a variable displacement means 68 , tar.κ line 7 , and pressure line 78 which is connected to line L5- Thβ hydraulic cylinder 1 shown on Pig. is replaced by the rotational hydraulic motor 1 which is loaded by a load 96 representin -trmaeβ of thβ motor vehicle. Thβ fly-wheel 9^ is attached to thβ common shaft 72 connecting pump 58, motor 66, and thβ primary motor 100 of the motor vehicle. Thβ variable displacement means 68 is modulated by thβ exhaust line prββaurβ dr feedback signal, which is equal P0«— ?« *«* « ■βasu Hv-fJ|I,linβ ,5 (through line 76) and thβ related signal line 3L5« Thβ exhaust line pressure drop feedback control system Including the variable displacement motor 66 , regulates the exhaust fluid pressure drop PQ*~— P« across spool valve 2 by varying thβ counterpressure *« =- *0« "~" ^'5 *** the exhaust power line L5 by thβ variable displacement means 68. Xa a simple ease, ene motor position command signal X being varied with the constant speed, will fβnβratβ a relatively constant velocity of motor 1 and the positional lag Δl proportional to this velocity. In general, thβ shaft vβlooity of motor 1 can be controlled by the speed of varying thβ motor position command signal XQ. During thβ deceleration of thβ motor vehicle, thβ kinetic energy accumulated by' a mass of thβ motor vehicle (load 96) is transmitted through motor 66 to thβ fly-wheel 9^. During the following acceleration of thβ motor vehicle,thβ kinetic energy accumulated by fly-wheel 9 iβ transmitted back through pump 58 to thβ motor vehicle. The exchange of kinetic energy between thβ motor vehicle (load 96) and thβ flywheel
9 is correlated with thβ fly-wheel speed fluctuations. It is assumed that a speed-torque characteristic of thβ Primary motor 100^ such as the electrical motor or the internal-combustion engine^ is soft enough to allow these fly-wheel speed fluctuations. The load adaptive, position feedback controlled, variable speed drive system having an independent regenerating circuitry is shown on Fig.11, which can be considered as thβ further development ( or modification ) of Pig.10.
In this drive system, a variable speed primary motor 92 of thβ motor vehicle is not connected to shaft 72 - the primary but is driving-y- shaft 98 of a variable displacemenf pump 90.
The tank line 38 of pump 90 iβ connected to tank 62.
The pressure line 5^ of pump 90 is connected through check valve 40 to the supply power line L2.
The variable speed primary motor 92, thβ related speed control block 92 included system.. The variable speed primary motor 92 is modulated by thβVsupply line pressure drop feedback signal ?2 — p , which is measured between line 5 (line 91) and line S12. primary
As a result, tKβVβupply line pressure drop feedbackori.mgrv control system is capable of maintaining thβ supply fluid pressure drop P, — Pn- across spool valve 2 by varying thβ Drimarv ^ ^pressure rate ?2 = PQ2 -+- ΔP2 in thβ supply power line ζ by varying thβ speed of the variable speed primary motor 92, such as the internal-combustion engine or the electrical motor.
On the other hand, the pump 58, shown on Fig.10 is replaced on Fig.11 by an assisting variable displacement pump 55 having an assisting variable displacement means 57 to make up an assisting supply line pressure drop feedbac N system. The line 36 of pump 55 is connected to tank 62. Thβ pressure line 30 of pump 5 is connected through check valve 44 to line I»2. The assisting ,. variable displacement means 57 iβ modulated by an assisting supply line pressure drop feedback signal P^ — ?Q2 , which is mea^re^^- line 30 ( through line 32 ) and line SL2. As a result, thβ assisting supply line pressure drop feedback control system is capable of maintaining thβ assisting supply fluid pressure drop P2R— Q2 acroβe spool valve 2 by varying thβ assisting pressure rate P2R=r PQ2 -f— Δ?2R in thβ supply power line 30. During thβ operation, the supply power line L2 is switched over to line 5 or line 30. whichever has thβ higher pressure rate, by thβ logic of chβcfc valves 40 and 44. assisting
TRβVprβssure drop command signal Δ 2λ is selected to be Just slightly larger than enβVprββsure drop command signal Δ ?2 ' Accordingly, while the speed of flywheel is still relatively high, mttiMb* P2Rβ P02 ^- ΔP2H will .χffi*$ $ffir.
?2=S ? QZ ~+* ΔP2 and, hence, thβ supply power line L2 will be connected to line 30 through check valve 44. t any other time, thβ supply power line 2 is connected to line 5- through check valve 40. In other words, thβ independent regenerating circuitry, including motor 66t pump 5, and fly-wheel 94, is piven a priority in supplying the fluid energy to the supply power line L2. This independent regenerating circuitry is automatically entering into, and is automatically withdrawing from thβ regulation of fhβVsuppi fluid pressure drop across spool valve 2* Thβ exchange of kinetio energy between the motor vehicle (load 6) and thβ fly-wheel 4 i« basically accomplished as considered above (for thβ schematic shown on Pig.10) t however, thβ undesirable interference between thβ primary motor 92, such as thβ electrical motor or thβ internal-combustion engine, and the regenerating circuitry is now eliminated.
It should be noted that the variable delivery means 93 of pump 90 can be employed for achieving some additional control objectives, such as maximizing thβ energy efficiency of the internal-combustion engine 92.
In fact, these additional control objectives can be similar to those which are usually persuaded in regulating the standart automotive transmissions of motor vehicles.
It should also be noted that schematic shown on Fig.11 is of a very general nature and can be further modified and (or) simplified. If there is no additional control objectives, such as just indicated, the variable speed primary motor 92 is replaced by a relatively constant variable for maintaining This case is illustrated by Fig.12 which is a modification of Fig.11 for the hydraulic press type application. In this case, the rotational hydraulic motor 1 is replaced by the double-acting cylinder 1. The exhaust line pressure drop feedback control system including motor 66 is adapted to maintain pressure P^ = PQ^ — Δp^ in the exhaust power line L3«
The potential energy of the hydraulic fluid compressed in chamber 10 of cylinder 1 is regenerated now by the independent regenerating circuitry through the exhaust power line L3 and thβ related exhaust line pressure drop feedback control system including motor 66. In fact, the schematic of Fig.12 i= easily understood just by comparison with Fig.il and Fig.9. For simplicity, the additional fluid power supply 50 is not shσwn on Fig.12.
Some preliminary generalization.
The motor load and the motor load means are the structural components of any energy regeneratingT^acTaptive fluid motor control system. For this reason, Fig.12 (as well as Fig.9 ) also shows the frame 190 ( of a hydraulic press 192 )} against which the chamber 10 of cylinder 1 is loaded. The compressed ÷bluid energy is basically stored within chamber 10 of cylinder 1; however, the stretching of frame 190 of press 192 may substantially contribute to the calculations of the over-all press energy accumulated under the load. It is noted that word "LOAD" within block 96 (see figures 10, 11, l6 to 22, and 26) is also considered to be a substitute for the words "the motor load means" and is ^ to all ihe possible applications of this invention. in a case of motor vehicle applica ions, the motor load means include a mass of a ".wheeled" motor vehicle ^as it is specifically indicated on the schematic of Fιg.22 ,
In the energy regenerating, load adaptive fluid motor control systems, such as shown on figures 9 to 12, it is often justified to consider the fluid motor and load means-as an integrated component. The fluid motor and load means include the fluid motor means anαVmo όr load means and accumulate a load related energy, such as a kinetic energy of a load mass or a compressed fluid energy of the fluid understood as motor-cylinder. The "exhaust fluid energy" isVa measure of the load related energy being transmitted through the exhaust power line (that is line L3 or line L5) • The "exhaust fluid energy" can also be referred to as the "waste fluid energy',' that is the energy which would be wasted unless regenerated.
There are basically two types of counterpressure varying means: a) the counterpressure varying means which are not equipped for recupturing the load related energy ( such as the exhaust line pressure drop regulator — see figures 1, 4, and 6 ) , and b) the counterpressure varying means which are equipped for recupturing the load related energy ( such as the exhaust line variable displacement motor — see figures Q, 10, 11, and 12). This counterpressure varying and energy recupturing means can also be referred to as the exhaust line energy recupturing means.
StUll other modifications of the exhaust line energy recupturing means will be considered later. Accordingly, there are basically two types of the load adapt ve fluid motor control systems : p ) the load adapt ve f.luid motor control systems which are not equipped tor regenera ing the load related energy ( see. rigures 1, 'l , and 6 ), and
1)) the load adaptive flu d motor control systems having an energy regenerating circuitry for regenerating the load related energy ( see figures 9 to 12 ). This second type of load adaptive fluid motor control systems can also be referred to as the regenerative adaptive fluid motor control systems. Still other modifications of the regenerative adaptive ' 3 id motor control systems will be considered later.
It should be noted that regenerative adaptive fluid motor control schematics being considered are the concept illustrating schematics only and, therefore, are basically free from the details, which are more relevent to the engineering development of the^?e concepts for specific applications. For example, the maximum and minimum pressures in hydraulic power lines must be restricted. Some design related considerations are summarized at the end of this .description.
General criterion of dynamic stability of combined component systems .
The load adaptive fluid motor position feedback control system control/' is typically a combination of at least three component feedback^ systems - the fluid motor position feedback control system, at ieast one exhaust line pressure drop feedback control system, and at least one supply line pressure drop feedback control system. In order to prevent a possible τmhπbontiial C fn -CQ,χ .ά . interference between the combj ned components systems, the pressure drop feedback control systems must be properly regulated both with respect to the fluid motor position feedback control system and with respect to each other.
Accordingly, a general criterion of dynamic stability of combined component systems (which are stable while separated) can be introduced by a set ol provisions (of by a combination of concepts) as follows:
( 1 ) preventing a substantial schematic operation interference between the pressure drop feedback control systems and the fluid motor position feedback control system ( this concept has been already discussed before);
(2) providing a significant dynamic performance superiority for the pressure drop leedback control systems against the fluid motor posi ion feedback control system, in order to prevent a substantial dynamic operation interference between the pressure drop feedback control systems and the fluid motor position feedback control system (this concept will be discussed later);
( 3 ) preventing a substantial pressure drop regulation interference between the supply and exhaust line pressure drop feedback control systems - this concept is discussed below.
The concept of preventing a substantial pressure drop regulation interference.
it should be noted that pressure-compensated flow characteristics which are shown on figures .3-D, 5-B, and 7-B, can generally be reduced to each of two asymptotic characteristics as follows _
(a) a motor static speed characteristic describing thβ hydraulic motor speed versus' thβ valve spool displacement, under the assumption that thβ hydraulic fluid is not compressible •
(b) a compression — decompression speed versus thβ valve spool displacement,, under thβ assumption that thβ hydraulic motor speed is equal to rβrσ.
Λs a result, the speed control of fluid motor 1 by any pressure drop feedback control system is generally effecied by the processes of compression-decompression of hydraulic fluid and, therefore, is substantially inaccurate. This speed control is, of course, still further effected by some other factors, such as the static and dynamic errows in maintaining the pressure drop.
It is also understood that a simultaneous speed control of fluid motor 1 by the supply and pressure drop feedback control systems may create avpressure drop regulation interference between these two systems. This pressure drop regulation interference may reveal itself in generating excessive pressure waves, producing hydraulic shocks, cavitating thβ hydraulic fluid, and accumulating an air in the hydraulic tracts.
Moreover, the pressure drop regulation interference may lead IO the over-all dynamic instability of the load adaptive fluid motor control system, such as the regenerative adaptive fluid motor control system.
The destructive conditions of pressure drop regulation interference can be avoided simply by preventing a simultaneous speed control of fluid motor 1 by two pressure drop feedback control systems, that is by the supply and exhaust line pressure drop feedback control systems. Without the loss of generality , the concept of preventing a pressure drop regulation interference is considered further more specifically for two examplifiβd groops of schematics as follows J
(a) the load adaptive scnematics having only one loadable chamber and, therefore, having only one pressure drop feedback control system controlling the speed of motor 1 at any given time - see figures 1, 4, 9, and 12;
(b) the load adaptive schematics having two loadable chambers, and therefore, having two pressure drop leedback control systems which potentially may participate simul aneously in controlling the speed of motor 1 - see figures 11 and 16 to 22. In the first groop of load adaptive schematics, the supply and exhaust line pressure drop feedback control systems will obviously never interfere. in these schematics, the speed of motor 1 is usually controlled only by a supply line pressure drop feedback control system ( that is by the primary supply line pressure drop feedback control system or by the assisting supply line pressure drop feedback control system ). The motor-cylinder 1 having only one loadable chamber is assumed to be loaded in only one direction by a static force. Accordingly, the motor load is measured by the pressure signals lιe exlιaust li e pressure drop feedback control system is usually in operation only during the decompression of chamber 10 oi motor i.
In the second groop of load adaptive schematics, a simultaneous speed control of motor 1 by the supply and exhaust line pressure drop feedback control systems is prevented by controlling the sequence of operation of these systems by the motor load of of motor 1, provided that pressure drop command signals ΔP-, , P'2R ' and ^p5 are selected so that:
-ΔP5 -_Δ.P2R>.ΔP2 • (3)
Let's consider now more specifically the second groop of load adaptive schematics. The magnitude and direction oi the motor load is conveniently measured by the pressure signals PQ-^ nnd P05- which are implemented for controiling the supply •:ind exhaust line pressure drop leedback control systems, respectively. The load pressure signals J-'n d,ld p05 are aiso used for controlling the sequence of operation of these pressure drop feedback control systems, as it is illustrated below.
Let's assume that wheeled vehicle is tested in a horizontal direction only. And let's consider briefly the related stop-and-go energy regenerating circle ( which is still further studied later - see Fig. b ).
1. The wheeled vehicle is moving with a constant speed. In this case, the motor load is positive, the load pressure signal PQ2 is relatively large, and the primary supply line pressure drop feedback control system is activated to maintain the primary supply fluid pressure drop P2 — PQ2 =^ ^^2 across spool valve 2. On the other hand, the pressure signal P05 is very small, and therefore, the exhaust line pressure drop feedback control system is not activated to maintain the exliaust fluid pressure drop PQg — P. r-r s across spool vaive 2. . Note that in this case, the exhaust fluid pressure drop P 5 — Pr is equal approximately to the primary supply line pressure drop command signal Z_ o • Provlclecl that supply and exhaust openings of valve 2 are identical. Note also that
2. The Wheeled vehicle is decelerated.
In this case, the motor load is negative, the load pressure signal P05 is large, and the exhaust line pressure drop feedback control system is activated to maitain the exhaust fluid pressure drop PQ5 — across spool valve 2. On the other hand, the pressure PQ2 is very small and has a tendency of dropping "below zero". In practical applications, a vacuum in motor line LI must be prevented by introducing a check valve (such as check valve 155 on figures 20 and 22) connecting line LI with the oil tank 62 ( or with a low-pressure hydraulic accumulator). Note that by virtue of expression (3), the process of deceleration should be started onle after this check valve is open. It is understood that in this setuation, the supply line pressure drop feedback control systems have no effect on the process of deceleration of motor 1.
3. The wheeled vehicle is completely stopped. In this case, the fluid motor 1 is not regulated.
4. The wheeled vehicle is accelerated.
In this case, the motor load is positive, the load pressure signal PQ2 is large, and the assisting supply line pressure drop feedback control system is activated to maintain the assisting supply fluid pressure drop P2R— P02 =rΔP2R acr ss spool valve 2. On the other hand, the pressure signal P05 is very small, and therefore, the exhaust line pressure drop feedback control system is not activated to maintain the exhaust fluid pressure drop PQ5 — P.. =r^P_ across spool valve 2. Note that in this case, the exhaust fluid pressure drop P 5 — Pc, is equal approximately to the assisting supply line pressure drop command signal _χ 2R- provided that supply and exhaust openings of valve 2 are identical . Note also that if P5 = 0 : P05=^P2R<^ 5 •
Finally, it can be concluded that in the load adaptive fluid motor control systems, the functions of the motor load are not limited to controlling separately each of the pressure drop feedback control systems. Indeed, the functions of the motor load are generally extended to include also the control over the sequence of operation of the supply and exhaust line pressure drop feedback control systems , in order to prevent a possible pressure drop regulation interference between these pressure drop feedback control systems.
The concept of providing a significant dynamic performance superiority.
It is important to stress that the concept of providin a significant dynamic performance superiority for the pressure drop feedback control systems against the fluid r.ctor position feedback control system is an integral part this invention. This concept introduces a criterion 21 dynamic stability of combined component systems which are stable while separated (provided that the concept of preventing a schematic operation interference and the concept of preventing a pressure drop regulation interference are already properly implemented). Λs it is already mentioned above, the load adaptive fluid motor position feedback control system is typically a combination of at least three component feedback control systems - the fluid motor position feedback control system, at least one exhaust line pressure drop feedback control system, and at least o e supply line pressure drop feedback control system.
The theory and design of the separate closed-loop systems are described in numerous prior art publications — see, for example, the books already named above, and also . a) Shinners S. M. , "Modern Control System Theory and ApplicationM, Second Edition. -Reading, Massachusetts : Addison- esley Publishing Company, 1972. b) Davis S. A. , "Feedback and Control System". New York . Simon and Shuster, 1 7**-.
It is further assumed that each of the separate component systems is linearized and, thereby, is basically described by the ordinary linear differential equations with constant coefficients, as it is usually done in the engineering calculations of electrohydraulic, hydromechanical, and hydraulic closed-loop systems. Note that the fluid motor position feedback control system (separated from other component systems) is especially easy to linearized if to admit that thβ expected regulation of the exhaust and supply fluid pressure drops is already "in place".
Let's consider (without thβ loss of generality) the load adaptive fluid motor position feedback δontrol system incorporating only three component systems — the fluid motor position feedback control system, only one exhaust line pressure drop feedback control system, and only one supply line pressure drop feedback control- system. In this case, the criterion of dynamic stability of combined component systems can be reduced to only five conditions as follows : (1) providing a dynamic stability of the fluid motcr position feedback control system ;
(2) providing a dynamic stability of the exhaust line pressure drop feedback control system ,•
(3) providing a dynamic stability of the supply line pressure drop feedback control system ;
( ) preventing a substantial dynamic operation interference between the exhaust fluid pressure drop regulation and the motor position regulation by providing a significant dynamic performance superiority for the exhaust line pressure drop feedback control system against the fluid motor position feedback control system ;
(5) preventing a substantial dynamic operation interference between the supply fluid pressure drop regulation and the motor position regulation by providing a significant dynamic performance superiority for the supply line pressure drop feedback control system against the fluid motor position feedback control system.
The presented above— first, second, and third, conditions of dynamic stability are the requirements to the separate component systems. The fourth and fifth conditions of dynamic stability define limitations which must be imposed on the separate component systems in order to combine them together. The design of the separate closed-loop systems for the dynamic stability and required performance is well known in the art, as already emphasized above. For this reason, it is further assumed, for simplicity, that the first three conditions of dynamic stability are always satisfied if the last two- conditions of dynamic stability are satisfied.
Because the last two conditions of dynamic stability are similar, they can also be specified by a general form as follows : preventing a substantial dynamic operation interference between the pressure drop regulation ( the exhaust or supply fluid pressure drop regulation) and the motor position regulation by providing a significant dynamic performance superiority for the pressure drop feedback control system ( the exhaust or supply line pressure drop feedback control system, respectively ) against the motor position feedback control system.
The provision of preventing "a substantial dynamic operation interference" is associated with the concept of providing "a significant dynamic performance superiority". The term "a substantial dynamic operation interference" is introduced to characterize the dynamic instability of combined component systems which are stable while separated. This dynamic instability can be detected in a frequency domain or in a time domain by ., M
= 1 (i )
Gύnp
or by =r 1, (5) kfd
respectively, where t
(DRp and aarree tthhee resonant frequency and thβ final
*fP transient time (respectively) of thβ fluid motor position feedback control system i jLj and t-d are thβ resonant frequency and the final transient time (respectively) of thβ pressure drop feedback control system.
Thβ closed-loop resonant frequency & n ( h& is ^R or C rf ) a located by a resonant peak of thβ closed-loop frequency-response characteristic and, therefore, is also often called "a peaking frequency". This resonant peak is typically observed on a plot of the amplitude portion of the closed-loop frequency-response characteristic. However, the resonant peak is observed. only if the system is underdQ ped.
For this reason and for simplicity, the appropriate
approximations of thβ ratio — c n also be
employed. For example, thβ possible approximation is
( 6 ) P 61
where J
( ) and J are thβ closed-loop bandwidths bp ^^bd for the position feedback control system and the pressure drop feedback control system, respectively.
Moreover, as thβ first approach (roughly approximately) .
where t
^oep and ^Cd ""• the open-loop cross-over frequencies for thβ position feedback control system and thβ pressure drop feedback control system respectively.
The final transient time f ( that is ^. or tfd ) of »e closed-loop system is tht total output-response time to the step input. The transient time tf is also often called "a settling- time" and is measured between t .= 0 and t s t^. — when the response is almost completed. The method of defining the closed-loop resonant frequency t| . > tnβ closed-loop bandwidth COL
, the open-loop cross-over frequency Ooc ' and the closed-loop final transient time - are well known in the art — see, for example, the above named books of S. M. Shinners , S. A. Davis , and A.F. D'Souza .
In accordance with equations (4-) and (5), there are two interrelated but still different aspects of dynamic instability of combined component systems which are stable while separated. Indeed, the .equation ( ) symbolizes a frequency resonance type phenomenon between thβ component systems. On thβ other hand, the equation (5) represents a phenomenon which can be viewed as an operational break-down of the combined component systems. Note that the exhaust and supply line pressure drop feedback control systems are the add-on futures and may fulfill their destination within the load adaptive fluid motor position feedback control system only iϊ *• the destructive impacts of "a substantial dynamic operation interference" are prevented by "a significant dynamic performance superiority?
Now, it is understood that if "a substantial dynamic operation interference" is identified by b) or (5), then "a significant dynamic performance superiority" should be identified by .
t , « and fp
> S ( 9)
' fd where.
• T is the minimum stability margin in a frequency domain,
S^ is the minimum stability margin in a time domain.
These minimum allowable stability margins can be specified approximately as ι ==r 10 and , =10. O t
The formulas (8) and (9) must be introduced into the design of tho load adaptive fluid motor position feedback control system. The way to do this is to design he separate component systems for the dynamic stability and required performance while the inequalities (3) and (9) for the combined component systems are satisfied.
The approximate connections between the resonant frequencies ar.d some other typical frequencies have been already illustrated by equations (6) and (7)«
While the equations (8) and (9) are valid for the second— and higher—order differential equations, the principal relationship between the final transient time ' t. and the resonant freαuency C „ is more easy to illustrate for the second-order equation
and
where ι y and Z " . are the input and output, respectively ι
the undamped natural frequency _Λ f
the damping coefficient *? = ) the dimensionless time 'c = CU& t . For this second-order equation, the output responses Z (c J to a unit step input ( while the initial conditions are zero) for various values of ? are well known in thβ art — see, for example, the above named books of S. M. Shinners and S. A. Davis.
Note that for the second-order equation fe~- ■ and, hence,
the inal • transient time r ~ — £- •
The final transient Λ dimensionless time ζ is a function of thβ damping coefficient ^T . More generally, when thβ right part of the second—order equation is more complicated, thβ final transient dimensionless time "S"^ is also effected by thβ right part of this equation.
In thβ case of using seoond-ordβr systems,
the ratio &^-Rd can be approximated by the rati.o &Λd and therefore t
where .
CO 2χι and ^ . are the undamped natural frequency and the f inal transient dimensionless time , respectively , for the position feedback control system j
(ι) and ^~fd arβ thβ undt*mPβd natural frequency and and the final transient dimensionless time, respectively, for the pressure drop eedback control system.
In general, for- the second— and higher—order systems, it can be still stated, by' the analogy with the second-
order system, that the ratio is basically dependent on the ratio • and is further dependent on thβ secondary factors, such as thβ effects of damping.
It is to say that expression (8) can be viewed as a basic
(or main) test on the dynamic stability of combined components systems which are stable while separated.
This main test is needed to prevent thβ frequency resonance type phenomenon between the component systems.
However, an additional test — equation (9) is still needed to prevent the operational break-down of the combined component systems.
In short, for thβ second- and higher-order systems i a) the expression (8) — alone is a necessary criterion for thβ dynamic stability of combined component rystems which arβ stable while separated > b) thβ expressions (8) and (9) — together are a sufficient criterion for the dynamic stability of combined component systems which arβ stable while separated .
Of course, still other terms, interpretations, and measure¬ ments can be generally found to further characterize what have been just clearly defined - based on thβ phy.ical considerations - as being "a substantial dynamic operation interference" and "a significant dynamic performance superiority. Adaptive fluid position feedback control: the scope of expected applications.
The load adaptive fluid position servomechanisms make it possible to substantially improve the energy, performance, and environmental characteristics of the position feedback control in comparison with the conventional fluid position servomechanisms. in particular, the load adaptive fluid position servomechanisms may combine the high energy-efficient and quiet operation with the relatively high speed and accuracy of performance. The artificial load adaptability of load adaptive fluid position servomechanisms is achieved by regulating the exhaust and supply fluid pressure drops by the exhaust and supply line pressure drop feedback control systems, respectively.
Because the artificial load adaptability is implemented by relatively simple design means, the load adaptive fluid position servomechanisms combine thβ very best qualities of thβ conventional fluid motor position feedback control systems and thβ naturally load adaptive, electric motor position feedback control systems. Moreover, the load adaptive fluid position servomechanisms may incorporate the energy regenerating circuitry.
Furthermore, maintaining the exhaust and supply fluid pressure drops across thβ directional control valve may protect thβ position closed-loop against such destructive conditions as generating excessive pressure waves, producing hydraulic shocks, cavitating thβ hydraulic fluid, and accumulating an air in the hydraulic tracts.
In other words, thβ transition to thβ adaptive servomechanisms makes it easy to control the fluid conditions in the hydraulic tracts and to provide a' "full hermetization" of thβ hydraulic motor.
Accordingly, the scope of potential applications of tho adaptive hydraulic position servomechanisms being considered is extremely wide. So, it is expected that thβ 4y
conventional hydraulic ( electrohydraulic or hydromechanical) position servomechanisms will be replaced almost everywhere by the load adaptive hydraulic position servomechanisms. It is.ilsc expected that many naturally load adaptive, electric motor position feedback control systems will also be replaced by thβ artificially load adaptive, hydraulic motor position feedback control sys'tβms.
In addition, it is expected that many βiβctrohydraulic, hydromechanical, nd electromechanical open-loop position control systems will also be replaced by the" load adaptive electrohydraulic and hydromechanical position servomechanisms .
The load adaptive fluid motor position feedback control systems can be used in machine tools (including presses), construction machinery, agricultural machinery, robots, land motor vehicles, ships, aircrafts, and so on.
In general, the load adaptive fluid position servomechanism can be viewed as a combination of a primary motor, such as the electrical motor or the combustion engine, and the load adaptive, position feedback controlled fluid power transmission, transmitting the mechanical power from a shaft of the primary motor to the load. The fundamental structural improvement of the position feedback controlled fluid power transmissions, as described in this invention, makes it possible to substantially increase the scope and the scale of their justifiable applications .
For example, the schematics shown on figures 9 and 12 can be used for constructing the high energy-ef icient hydraulic presses. The load adaptive hydraulic press may have advantages against the conventional hydraulic and mechanical presses due to a combination of factors as follows: i. ThθVeriergy-efficiθncy of the hydraulic system combining the load adaptive primary power supply and the load adaptive regeneration of energy.
2. Superior performance and environmental characteristics including: the smooth and quiet operation of the moving slide, the smooth compression and decompression of the hydraulic fluid, the high speed, accuracy, and dynamic performance potentials.
3. The press is easy to control with respect to the moving slide position, stroke, speed, and acceleration. The press maximum tonage is also easy to restrict for the die-tool protection.
4. Simplicity of design - only one regenerative adaptive hydraulic position servomechanism is required to provide all the benefits described.
Finally, it should be noted that schematics shown on figures 4 and 12, make it possible to absorb the shocks generated by a sudden disappearance of load, for example, during the punching operations on hydraulic presses. This is accomplished by decelerating thβ motor-cylinder 1 just before the load disappears to provide thβ valve spool to be close to its neutral point ( X ss 0 ) . Just a ter thβ load disappears, thβ position feedback control systam locks thβ fluid in chamber 11 or even connects this fluid with thβ supply power line 12. It means that thβ potential energy of thβ fluid compressed in chamber 10, is used mostly to compress thβ fluid in chamber 11 and, finally, is converted to a heat . <
Adaptive fluid motor feedback control.
Fig.13 shows a generalized model of the load adaptive fluid motor output feedback control systems which include an independent energy regenerating circuitry. This model can be viewed as a further development of Fig.8 in view of figures 11 and 12 and is mostly self-explanatory. Note that the position feedoack control means ( block H- ) and the related signals Xχ , XQ , and X , which are shown on Fig.8,* are replaced by the ( motor ) output feedback control means ( block -M ) and the related signals M-j_ , MQ, and ΔM, which pre shown on Fιg.l3>
More specifically, the motor position X, , the position input-command signal X , and the position feedback control error signal ΔX are replaced by their "generic equivalents" — the motor output M, , the related input-command signal MQ , and tho motor output feedback control error signal ΔM, respectively.
By the analogy with the load adaptive fluid motor position feedback control system, the motor output feedback control error signal ΔM is produced by the output feedback control means (block '- ) in accordance with a difference between the input-command signal H and the motor output M, .
Clearly, the motor output is a generic name at least for the motor position, the motor velocity, and the motor acceleration. Accordingly, the load adaptive fluid motor output feedback control system is a generic name at least for the following systems . a) the load adaptive fluid motor position feedback control system. b) the load adaptive fluid motor velocity feedback control system? c) the load adaptive fluid motor acceleration feedback control system.
The general criterion of dynamic stability of combined component systems, which was formulated above with respect to the load adaptive fluid motor position feedback control system, is also applicable to the load adaptive fluid motor output feedback control system. In particular, the concept of providing "a significant dynamic performance superiority", which ■ϊτ ϊ-formulated above with respect to the load adaptive fluid motor position feedback control system, is also applicable to the load adaptive fluid motor output feedback control system.
regenerative A generalized model of the ^ adaptive fluid velocity servomechanisms is shown on Fig.l . This model is derived from the one shown on Fig.13 just by replacing the
(motor) output feedback control means (block 4-- ) and the related signals M, , M , and ΔM by the velocity feedback control means ( block -V ) and the related signals V. , V , and Δv, respectively. It is to say that the schematics for the adaptive fluid velocity servomechanisms being considered can also be derived from the above presented schematics for the adaptive fluid position servomechanisms just by replacing the position feedback control means (block ) and the related signals X, , X , and
ΔX by the velocity feedback control means (block -V ) and the related signal V,, V , and Δv, respectively.
The motor velocity V- is the velocity of the moving part 21 of the fluid motor 1. In fact, the motor velocity V. can also be viewed as a mechanical signal — the output velocity signal of the load adaptive fluid motor velocity feedback control system. The motor velocity V, is measured by the velocity sensor, which is included into block 4-V and is connected to the moving part 21 of the fluid motor 1. The velocity feedback control error signal Δv is produced by the velocity feedback control means (block k-V) in accordance with a difference between the velocity input- command signal V and the motor velocity V . It is reminded that at the balance of the position feedback control j ΔX =0 and the spool of valve 2 is in the neutral spool position for any given value of the position command signal X . Accordingly, at the balance of the velocity feedback control: Δv = Oj however, the spool of valve 2 is not generally in the neutral spool position but is in the position which corresponds to the given value of the velocity command signal V . It is already understood that the velocity feedback control means (block *4~V ) can be still further described basically by the analogy with the above brief description of the position feedback control means (block )• The optional physical structure of the velocity feedback control means ( block ^-V ) is also disclosed ' by numerous prior art patents and publications describing the conventional fluid motor velocity feedback control systems and the related velocity feedback control technique — see, for example the books already named above.
The schematic shown on Fig.16 can be used for constructing the load adaptive , velocity feedback controlled, fluid power drive systems for the motor vehicles. This schematic is derived from the one shown on Fig.11 by replacing the position feedback control means (bloc* ) and the related signals X , X, , and ΔX by the velocity feedback control- means (block --V) and the related riτnals V , V, , and ΔV, respectively. In addition and for simplicity, the five-way spool valve 2 shown on Fig.11 is replaced by the four-way spool valve 2 shown on Fig.16. Accordincly, the supply power line Lό and the exhaust power line 3 are eliminated.
The four-way spool valve 2 is considered now to be a one- spool directional valve— it's^ aTTbe moved only down from the neutral spool position and can be returned back to the neutral spool position only ( which is shown on Fig.lό ). Regenerative adaptive fluid motor control.
A generalized model of the regenerative adaptive fluid motor open-loop control systems is presented by Fig.15 which is derived from
Fig.13 just by eliminating the output feedback control means (block - ) and the rel.ated signals M , M, , and Δ . The schematics for the load adaptive fluid motor open-loop control systems can be derived from the above presented schematics for thβ load adaptive fluid motor position feedback control systems just by eliminating thβ position feedback control means ( block ) and thβ related signal XQ, Xχ, and ΔX.
The open-loop schematic, which is shown on Fig.17 , is derived from the one shown on Fig.16 just by eliminating the velocity feedback control means ( block -V ) and the related signals V , V, , and ΔV.
The schematic of Fig.17 can be used for constructing the high energy-e ficient load adaptive motor vehicles, as it will be still further discussed later.
The general criterion of dynamic stability of combined component systems, which was formulated above with respect to the load adaptive fluid motor position feedback control systems, is also applicable to the load adaptive fluid motor open-loop control systems. in particular, the concept of providing "a significant dynamic performance superiority", which is formulated above with respect to the load adaptive fluid motor position feedback control system, is also applicable to the load adaptive fluid motor open-loop control system.
Λ significant dynamic performance superiority of any pressure drop feedback control system against the fluid motor open-loop control system can be established, for example, by providing basically a significantly larger closed-loop bandwidth for this pressure drop feedback control system in comparison with an open -loop cross-over frequency of the fluid motor open- loop control system-.
General principle of coordinated control : the constructive effect of motor load.
As it is already mentioned above, a regenerative adaptive fluid motor control system is typically a combination of at least three component control systems - a fluid motor control system, at least one exhaust line pressure drop feedback control system, and at least one supply line pressure drop feedback control system. The fluid motor control system may or may not include the output feedback control means.
Let's assume that for any given regenerative adaptive fluid motor control system:
(1) all the separate component systems are dynamically stable ( and provide the required dynamic performance ) and
(2) the general criterion of dynamic stability of combined component systems is satisfied, which means that:
(a) the concept of preventing a schematic operation interference, which was presented above, has been already properly implemented}
(b) the concept of providing a significant dynamic performance superiority, which was presented above, has been also properly implemented)
(c) the concept of preventing a pressure drop regulation interference, which was presented above, has been also properly implemented.
Under all these preconditions, one general principle can now be formulated, in order to clearly visualize why all the component systems will be working in unison to provide an operative regenerative system. This "general principle of coordinated control" can be formulated as follows:
In a regenerative adaptive fluid motor control syste the component systems will not interfere and will not "fall a part", but instead will be working in unison, to provide an operative regenerative system, by virtue of controlling all the pressure drop feedback control systems from only one "major coordinating center" — that is' by the only one (total) motor load. This general principle reveals the constructive effect of motor load.
in order to illustrate this principle more specifically, let's consider, for example, a regenerative adaptive fluid motor drive system for the motor vehicle. In accordance with figures 10, 11, 16, and 17, the magnitude and direction of motor load of motor 1 are conveniently measured by the pressure signals PQ2 a d Pø5- These pressure signals can also be viewed as the load related, input-command signals for the supply and exhaust line pressure drop feedback control systems, respectively. It means that all the pressure drop feedback control systems are, indeed, controlled in unison by the motor load of motor 1.
Finally, it can also be concluded that in the load adaptive motor vehicles, the vehicle speed is controlled by the driver via the fluid motor control system, while the energy supply and regeneration processes are all controlled in unison by the motor load via the pressure drop feedback control systems. In short, the load adaptive motor vehicle drive system is, indeed, an operative regenerative system having all the components working in unison. SOME EXΛMPLIFIED SYSTEMS
Adaptive fluid control and the motor vehicles. motor The load adaptiveVvehicle drive systems, like the one shown on Fig.17, mny have advantages against the conventional motor vehicle drive systems in terms of such critical characteristics as energy efficiency, environmental efficiency, reliability, controlability, and dynamic performance. Some of the underlying considerations are:
1. By virtue of the load adaptability, the task of controlling the speed of the motor vehicle is conveniently separated from the tasks of controlling the energy supply and conservation.
2. The primary supply fluid pressure drop regulation by the variable speed primary motor (engine) 92 has an effect of the energy supply regulation in accordance with the actual energy requirements. , ^, regulation
3. The exhaust fluid pressure drop^and thβ independent regenerating circuitry make it possible to create thβ schematic conditions, under which thβ energy accumulated during thβ deceleration of thβ motor vehicle is reused during thβ following acceleration of thβ motor vehicle. The energy accumulated during the vehicle down-hill motion will also be reused. . A the presence of load adaptive control, a standart braking system of the motor vehicle can be used mostly as a supplementary ( or emergency ) braking system.
5. In the load adaptive motor vehicles, a relatively smaller engine can usually be used.
6. Moreover, this smaller engine can be substituted by two still smaller engines, only one of which is operated all the time, while the second engine is switched-in only when needed - for example, when the vehicle is moving up-hill with a high speed, as it will be explained more specifically later. 7. The air pollution effect of the motor vehicles will be substantially reduced just by eliminating the waste of energy engines, and brakes-
8. In the load adaptive motor vehicles, no controlled mechanical transmission is needed.
9- The schematics of figures 11, 16, and 17 can be modified by replacing the variable speed primary motor 9 by the constant speed primary motor 100 and by using the variable displacement means 93 of pump 90 for regulating the supply fluid pressure drop p — Pø-irrΔP - as it was already illustrated by Fig.12.
Adaptive fluid control and the City Transit Buses.
The load adaptive drive system, such as shown on Fig.17, is especially effective in application to the buses which operate within the cities, where a stop-and-go traffic creates the untolerable waste of energy, as well as the untolerable level of air pollution. Let's assume , for simplicity, that the bus is moving in a horizontal direction only. And let's consider, for example, the process of bus deceleration - - acceleration beginning from the moment hen the bus is moving with some average constant speed and the "red light" is ahead. Up to this moment the spool of valve 2 have been hold pushed partially down by the driver so that this valve is partially open.
In the DΓOCRRS σf bus deceleration • the spool of valve 2 is being moved up- to clos? this valve; the pressure 0ς in line l is increasing. the pressure P- in lineL? is also increasing) the exhs'UP fluid energy of the exhaust fluid flow is bei-g transmitted through motor 66 to the fly-wheel accumulator
9^. '
As the spool valve 2 is finally closed, the bus is almost stopped pnd the complete stop is provided by using the bus brakes - as usually.
In the process of bus ?cceleration : the spool of valve 2 is being moved down - to open this valve, the pressure P0± in line LI is increasing! the Pressures Λ and ^ are also increasing, however P.R > ^ and therefore check valve S open, and check valve O is closed,- the energy accumulated by fly-wheel 9^ is transmitted through pump 55 , check valve , lines L2 and LI to the motor 1. When the energy accumulator 94 • is almost discharged, the pressure P^ ^ is being dropped so that the check valve is closed, and the check valve 40 is open permitting the engine 9 to supply the power flow to the fluid ^otor 1.
The load adaptive drive systems, like the one shown on Fig.17, can also be characterized by saying that these drive systems incorporate the energy regenerating brakes.
Adaptive fluid control with the hydraulic accumulator.
The regenerative adaptive fluid control schematic which is shown on Fig.18, can also be used for the motor vehicle applications, and in particular, for the buses which operate within the cities. This schematic will be studied by comparison with the one shown on Fig.17.
The fly—wheel 9^ shown on Fig.17 is substituted by a hydraulic accumulator 122 shown on Fig.18. Accordingly, the exhaust line variable displacement motor 66 is replaced by the exhaust line constant displacement motor 116 driving tne exhaust line variable displacement pump 120 which is powering the hydraulic accumulator 122 through check valve 136. oU
The exhaust line variable displacement pump 120 is provided with the variable displacement means 130 which is used to maintain counterpressure P- ;=• PQ-— Δp5 in Lhe exnaust power line L5 - as before. In other words, a counterpressure transformer including fluid motor 116, shaft 110, fluid pump 120, tank lines 74' and 134, and power lines 78 and 132, is implemented to make up the counterpressure varying and energy recupturing< means of the exhaust line pressure drop feedback control system maintaining counterpressure P. PQ5 —ΔP, in tne exhaust power line L5.
The assisting variable displacement pump 55 is replaced by the assisting constant displacement pump 114 being driven by the assisting variable displacement motor 118 which is powered by the hydraulic accumulator 122. The assisting variable displacement motor 118 is provided with the variable displacement means 128 which is used to maintain pressure P2R = P02_r* P2R in- the line 30 as before- I" other words, a pressure transformer, including fluid pump 114, shaft 112, fluid motor 118, tank lines 36 and 126, and power lines 30 and 124, is implemented to make up the assisting variable delivery fluid power supply of the assisting supply line pressure drop feedback control system maintaining pressure 2R — P02 "T;"-^AP2R ιn the line 3°_
Adaptive fluid control: the combined energy accumulating means.
It is understood that many other modification and variations of regenerative adaptive fluid control schematics are possible. These schematics may include the fly-wheel, the hydraulic accumulator, the electrical accumulator, or
_any combined energy accumulating means.
The examplified schematic showing the combined energy accumulating (and storing) means is presented by Fig.19 which is basically a repitition of Fig.18, however, two major components are added, the electrohydraulic energy converting means 142 and the electrical accumulator 144.
In addition, and just for diversity of the drawings presented , the variable speed primary motor 92 is replaced by the constant speed primary motor 100, so that now the variable displacement mechanism 93 of pump 90 is used for regulating the supply fluid pressure drop P? — , as it was already illustrated by Fig.12. As the hydraulic accumulator 122 is almost fully charged, an excess fluid is released from this accumulator, and a hydraulic energy of the excess fluid is converted through the electrohydraulic energy converting means 142 to the electrical energy of electrical accumulator 144.
On the other hand, as the hydraulic accumulator is almost fully discharged, the energy is transmitted back from the electrical accumulator 144 to the hydraulic accumulator 122.
The schematic of Fig.19 can be characterized by that the combined energy accumulating (and storing) means include the fluid energy accumulating means being implemented for powering the electrical energy accumulating means. More generally, the combined energy accumulating (and storing) means may include major (primary) energy accumulating means being implemented for powering supplementary (secondary) energy accumulating means.
Note that a common electrical power line can also be employed as an equivalent of the energy accumulating (and storing) means. For example, the combined energy accumulating (and storing) means may include fluid energy accumulating means (hydraulic accumulator 122 on Fig.19 ) being implemented for powering the electrical power line (replacing electrical accumulator 144 on Fig.19 ). In this case, the electrical power line will accept an excess energy from the hydraulic accumulator 122 and will return the energy back to the hydraulic accumulator 122 — when it is needed. Adaptive fluid control with a .variable displacement motor driving the load.
Fig.20 is basically a repetition of Fig.10; however, trie variable speed primary motor 9 is introduced now by the variable speed primary Internul-combussion engine 92.
In addition, tne c,
— - the constant displacement motor 1. driving load 96 is replaced by a variable displacement motor 15n driving the same load.
The variable displacement means 152 of motor 150 are constructed to make-up the displacement feedback control system including a variable displacement mechanism (of motor 150 ) and employing a displacement feedback control errow signal AD, is generated in accordance with a difference between a spool displacement ( command signal ) D of valve 2 and a mechanism displacement ( feedback signal ) D-^ of the variable displacement mechanism of motor 150. The displacement feedback control errow signal
Δϋ=D0 — D-^ is implemented for modulating the variable displacement mechanism of motor 150 for regulating the mechanis displacement D, of the variable displacement mechanism of motor 150 n accordance with the spool displacement Do of valve 2. It should be emphasized that the displacement feedback control system, which is well known in the art, is, in fact, the position feedback control system and that, therefore, the general position feedback control technique, which is characterised above with respect to the fluid motor position feedback control system, is also basically applicable to the displacement feedback control 'system.
Λs the spool of valve 2 is moving down from the" zero" position shown on Fig.20, there are two consecutive stages of speed regulation of motor 150 , the lower speed range is produced by changing the actual
(orifice ) opening of valve 2, the higher speed range is produced by changing the displacement of motor 150. Speaking more specifically , the lower speed range of motor 150 is defined between the "zero" spool position and the point of full actual ( orifice ) opening of valve 2. Up to this point, the command signal D is kept constant, so that the displacement of motor l^O is maximum and is not changed.
The higher speed range of motor 150 is located beyond the point of full actual ( orifice ) opening of valve 2. Beyond this point ( due to the spool shape of valve 2 ) the further spool displacements do not change any more the opening of valve 2. On the other hand, beyond this point, the oomm nα alg-r d. D is being reduced b,y tne i'urτner spool displacements of valve Z. Accordingly, tne displacement ___ j} — D of tne variable displacement mechanism of 1 o motor 150 is being also reduced by tne displacement feedback control system. The smaller the displacement of motor 150, the higher the speed of this motor ( and tne smaller the available torgue of this motor ) .
Fig. 0 also illustrates the use of check valves for restricting the maximum and minimum pressures in the hydraulic power lines. The check valve 154 is added to very efficiently restrict the maximum pressure in the exhaust motor line L4 by relieving an excess fluid from this line ( through check valve 15^ ) into the high-pressure hydraulic accumulator 122. The check valve 155 is added to effectively restrict the minimum pressure in the supply motor line Ll by connecting this line ( through check valve 155 ) with the tank 62. ote that tank 62 can generally be replaced by a low-pressure hydraulic accumulator ( accompanied by a small— supple etary tank ) . Adaptive fluid control with a regenerative braking pump.
In the motor vehicles, such as the City Transit Buses, the available braking torque should be usually substantially larger than the available accelerating torque.
Fig.21 is basically a repetition of Fig.18 ; however, the constant displacement motor 1 driving the load 96 is also driving a regenerative braking variable displacement pump 170 which is used to increase the available regenerative braking torque. The tank line 176 of pump 170 is connected to tank 62. The pressure line 178 of pump 170 is connected through check valve 17 to the hydraulic accumulator 122. The flow output of pump 170 is regulated in accordance with the pressure rate P^ in the motor line L4 conducting a motor fluid flow from the fluid motor 1, as it is more specifically explained below.
The variable displacement means 99 of pump 170 are constructed to make-up a displacement feedback control system including a variable displacement mechanism ( of pump 170 ) a displacement feedback control errow signal Δd. generated in accordance with a difference between a command-displacement signal d =c Pn r. ( where C is o p 05 P a constant coefficient ) and a mechanism displacement ( feedback signal ) d1 of the variable displacement mechanism of pump 170. Λ pressure-displacement transducer converting the pressure signal
P05 to ^ne Proportional command-displacement signal d — C *Pnt- is included into the variable displacement means 99 of pump 170. This transducer may incorporate, for example, a small spring-loaded hydraulic cylinder actuated by the pressure signal p °5* The displacement feedback control errow signal A~ = dQ— d1 is implemented for modulating the variable displacement mechanism of pump 170 for regulating the mechanism displacement άλ of the variable displacement mechanis of pump 170 in accordance with the command signal d . . . & o
(, and hence, in accordance with the pressure rate 6J
P = d /C in the motor line L ).
It should be emphasized that the displacement feedbabk control system, which is well known in the art, is, in fact, the position feedback control system and that, therefore, the general position feedback control technique, which is characterised above with respect to the fluid motor position feedback control system, is also basically applicable to the displacement feedback control system.
In general, the displacement feedback control circuitry of pump 170 is adjusted so that, while the pressure PQ e- in the motor line L4 is comparatively low, this circuitry is not operative and d, = 0. As the pressure PQ C. in the motor line iΛ is further raising-up, the displacement d.. of pump 170 is increasing acdordingly, so that the total regenerative braking torque is properly distributed between the fluid motor 1 and the regenerative braking pump 170.
Note that a significant dynamic performance superiority must be provided for the displacement feedback control system against the energy recupturing (recuperating) pressure drop feedback control system, in order to prevent their substantial dynamic operation interference. The concept of providing a "significant dynamic performance superiority" have been already generally introduced before and is further readily applicable to the displacement feedback control system versus the energy recupturing (recuperating) ^pressure drop feedback control system. Adaptive fluid control patterns.
Fig.22 is basically a repetition of F.ιg.20, however, the variable speed primary internal-combussion engine 92 is now replaced by a relatively constant speed primary internal-combussion engine 100, while tne varialtole displacement pump 90 is adapted now for maintaining the pressure P^ = PQ2 ---Δ P2 i line 5'J- » as it was already illustrated, for example, by Fig.19- In addition, the variable displacement motor 110 ynd the constant displacement pump 114 are replaced by the constant displacement motor 190 and the variable displacement pump 19>I, i n order to provide a wider range of regulation of pressure. 2R= P 02"*~^P 2R ιn line -0, The assisting constant displacement motor 198 Is powered by the hydraulic accumulator 122 (through shut-off valve 299 ) and is driving the assisting variable displacement pump 194 which is pumping the oil from tank 62 back into the accumulator 122 (through check valve 204 and shut-off valve 299 ) . Actually, the output flow rate of accumulator 122 (in line 210) is equal to a difference between the input flow rate of motor 198 (in line 200) and the output flow rate' of pump 194 (in line 1 4).
The exhaust from motor 190 is used to power the line 30. assisting , . he1 Variable displacement means 196 of pump 19'' is modulated by tπe'vf}Tes ire drop feedback signal P2R — P()2 bo maintain pressure P_ rP02, - P2R in line 30—
— as before. Time torque of pump 194 counterbalances the torque of motor 190. As the displacement of pump 19'^ is varied ( by the pressure drop feedback control syscem ) , the pressure P?R in line 30 can be regulated from "almost 7,ero" τ,o the ••maximum" , accordingly. The check valve 208 connects Line 132 ( of pump 120 ) with the tank 62.
The shut-off valve 29J? is controlled by the load pressure signal PQ2 • The check valve 208 and shut-off valve 299 are considered to be optional and are introduced only to illustrate more specifically some exemplified' patterns of controlling the load adaptive exchange of energy between the fluid mo Lor and load means and the energy accumulating means. The related explanations are presented below .
Let's consider, first, a simple case, when the motor vehicle is moving in a horizontal-direction only. While the motor vehicle is moving witn a constant speed ( or is being accelerated)^ the pressure PQ ) ( in line 1 ) is very small and does not effect the initial displacement of pump 120^ provided that pressure drop command signals P2. /\ P2R.and -^ P5 are selected so that ./\ P5 > \ P2R SΛP2' as it is required by expression (3).
This initial pump displacement is made just slightly negative, in order to provide for the pump 120 a very small initial output ( in line 13^ ) directed to the tank 62, and thereby, to provide for the exhaust fluid flow (in line L5 ) a free passage through motor 116 to the tank 62. In oi er words, while the pressure signal PQC s very small, the check valve 208 is open, the cneck valve 136 is closed, 3tnd the pump 120 is actually disconnected from the accumulator 122. fi e the motor vehicle is being decelerated, the displacement o£ pump 120 is positive, the check valve 200 is closed, the check valve 13 is open, and the kinetic energy of a vehicle mass is converted to tne accumulated energy of accumulator 122, as it •'already explained above.
In a general case, the motor vehicle is moving in a horizontal direction, up-hill, and down-hill, and with the different speeds, accelerations, and decelerations; however, all what counts for controlling the energy recupturing pressure drop feedback control system, is the load rate and direction ( which are measured by the pressure signals Pn,- and P02 )• While the pressure signal P& c. is very small, the pump 120 is actually disconnected from the accumulator 122, and the exnaust fluid flow is passing freely through motor 116 to the tank 62. As tne pressure signal P^ e is increasing, the kinetic energy of a vehicle mass is converted to the accumulated energy of accumulator 122.
On the other hand, all what countffor controlling the primary and assisting supply line pressure drop feedback control systems (and the shut-off valve 299), is also just the load rate and direction (which are measured by the load pressure signals PQ2 and PQ5 )• While the pressure signal PQ2 is very small, the shut-off valve 299 is closed. After the pressure signal PQ2 is measurably increased, the shut-off valve 299 is open.
In short, there are many regenerative adaptive fluid control patterns which are basically adaptive to a motor load, while are also responsive to the specific needs of particular- applications. All the variety of the regenerative adaptive fluid control patterns is^fact , within the scope of this invention. Fig.22 is still further studied later - with the help of supplementary figures 23 to 25.
Adaptive fluid control: two major modifications.
There are two major modifications of adaptive fluid control having an independent regenerating circuitry. The first major modification is identified by using the variable speed primary motor 92 for regulating the primary supply fluid pressure drop, as illustrated by figures 11, 16, 17, 18, 20, and 21. The second major modification is identified by using the variable displacement mechanism of the variable displacement primary pump 90 for regulating the primary supply fluid pressure drop, as illustrated by figures 12, 19, and 22. It is important to stress that these two major modificationε are often convertible. For example, the schematics shown on figures 11, 16 , 17, 18, 20, and 21 can be modified by replacing the variable speed primary motor 92 by a constant speed primary motor 100 and by using the variable displacement primary mechanism of pump 90 for regulating the 'supply fluid pressure drop P2 — * as it is illustrated by figures 12, 19, and 22.
The transition to the modified schematics is further simplified by providing a constant speed control system for the variable speed motor 92 and by converting, thereby, this variable speed motor to a constant speed motor.
Regenerative adaptive drive systems.
It πhould be emphasized that the combined schematics providing an automatic transition from the one mode of operation to the other are especially attractive for the motor vehicle applications. The examplified modi ications of combined sonematicπ can be briefly characterised as follows.
1. The motor vehicle is first accelerated by actuating tne variable displacement mechanism of pump 90 — as illustrated ana by Fig.227V"Ts further accelerated by actuating the variable speed primary interrial-oombussion engine — as illustrated by
Fig.20. This first modification of combined schematics can be viewed as a basic ( or first ) option of operation.
2. The motor vehicle is first accelerated .by actuating tne variable displacement mechanism of pump 90 — as illustrated by Fig. 2, is further accelerated by actuating the variable speed primary i nternai-oombus ion engine — as illustrated by Fig. 0, and is still further accelerated by actuating the variable displacement mechanism of motor 15 — as illustrated by figures 20 πiid 22. Note that in this case, the engine will be usually fully loaded only during the third stage of speed regulation — just after the displacement of motor 1 is sufficiently reduced. Note also that the minimum possible displacement of motor 150 must be restricted by the desirable maximum of engine load ( which can be measured, for example, by the desirable maximum of pressure Pft? in line Ll of motor 150 ).
3- The motor vehicle is first accelerated by actuating the variable displacement mechanism of pump 90 — as illustrated by Fig.22, and is further accelerated by actuating the variable speed primary internal-combussion engine — as illustrated by Fig. 0. Contrary to point 2, there is no third consecutive stage of speed regulation ( by using the variable displacement motor 150 ). Instead, the displacement of motor 150 is controlled independently by using the pressure signal PQ2 which is provided by line Ll. The larger the pressure signal Pn? , the larger the displacement of motor 150 — within the given limits, of course, 4. The motor vehicle is provided with two relatively small engines. The first engine time. The second engine is while the motor vehicle io moving up-hill with a high speed. Enc engine is driving * sep rate pump ( like pump 90 ). Each engine-pump instalation is working with a separate spool valve ( likn πpool vnlvo 2 ).
'j ■ Tho so oπd option of operation ( tipc point 2 ) is applied to the first engine-pump ins tala tion. of the two-engine vehicle of point ^«
6. The first option of operation ( rsee point ] ) is npplied to the πecond engine-pump nπ t nl t ion of the two-engine
7 . T h e th i rd op t i on o f o pera t i on ( n e e poin t 3 ) in a plie to the first engine-pump instalation of the two-engine vehicle of point 4.
8. The third option of operation ( see point ) is also applied to the second nngine-pu p .instalation of the two-engine vehicle of point '|,
9. The independent regenerating circuitry, such as shown on figures 11 to 22, can be easily switched-off by the driver in the process of operating a motor vehicle. This can be accomplished by using a directional valve switching over the exhaust power line L5 from the energy regenerating circuitry to the tank.
10. Note that regenerative adaptive drive system, such as shown on Fig.22, can be modified by replacing the "stationary" exhaust line energy recupturing means ( the constant displacement motor 116 driving the variable displacement pump 120) and the "stationary" assisting variable delivery fluid power supply
( the constant displacement motor 198 driving the variable displacement pump 194) by only one "shutle-type" motor-pump instalation including a constant diplace ent motor driving a variable displacement pump. Let's assume, for example, that wheeled vehicle is moving in a horizontal direction only. While the vehicle is decelerated, this motor-pump instalation is switched-in to perform as the "made-up" exhaust line energy recupturing means. While the vehicle is accelerated, this motor-pump instalation is switched-in to perform as the "made-up" assisting variable delivery fluid power supply.
Integrated drive system.
The energy regenerating, load adaptive drive system of a wheeled vehicle can be still further modified to provide an optional mechanical connection of the engine shaft with the wheels of the vehicle. This optional mechanical connection can be used, for example, for long-distance driving.
The de*sign of modified-integrated drive system may include an integrating mechanical transmission to select one of two alternative - component systems as follows:
1. The basic regenerative adaptive drive system - see figures 17 to 22. In this' case, the engine of a vehicle is connected with the primary pump 90. The back axil of α vehicle is driven by the constant displacement motor 1 (or by the variable displacement motor 150).
2. The optional conventional power train. In this case, the shaft of the engine is connected mechanically to the back axil of a vehicle.
CONCLUSIONS
Regenerative adaptive fluid motor control: . the energy recuperating pressure drop feedback control system.
A regenerative adaptive fluid motor control system having an independent regenerating circuitry (see figures 11 to 22 ) p ^ is an integrat*ng- system incorporating only two major components i * Q n) the two-way load adaptive fluid motor control system which is adaptive to the motor load along the exhaust and supply power lines of the- spool valve 2, and b) the two-way load adaptive energy regenerating system which is also adaptive to the motor load along the exliaust and supply power lines of the spool valve 2.
The regenerative system having an independent regenerating circuitry is charactirized by that the primary and assisting supply line pressure drop feedback control systems are separated. On the other hand, the exhaust line pressure drop feedback control system (which can „ also be referred to d the energy recupturing pressure drop reedback control system ) is shared between the two-way load adaptive fluid motor control system and the two-way load adaptive energy regenerating system. The energy recupturing pressure drop feedback control system includes an exhaust line energy recupturing means for varying a counterpressure rate in the exhaust power line and for recupturing a load related energy, such as a kinetic energy of a load mass or a compressed fluid energy of a fluid motor-cylinder. The energy recupturing pressure drop feedback control system and the exhaust line energy recupturing means can also be referred to as the energy recuperating pressure drop feedback control system and the exhaust line energy recuperating means, respectively.
Load adaptive energy regenerating system.
The above brief description of examplified load adaptive energy regenerating systems ( .see figures 11 to 22 ) can be still further generalized and extended by the comments as follows.
l ti d t
2. The fluid motor and load means include the fluid motor means and the motor load means and accumulate a load related energy ( such as a kinetic energy of a load mass or a compressed fluid energy of the fluid motor-cylinder ) for storing and subsequent regeneration of this load related energy.
3. Λs it was already mentioned before, the "exhaust fluid energy" of the exhaust fluid flow is understood as a measure of the load related energy being transmitted through the exhaust power line (that is line L3 or line L5). The "exhaust fluid energy" can also be referred to as a "waste fluid energy", that is the energy which would be wasted unless regenerated .
4. The f rs e'nergy converting means include the energy recupturing pressure drop feedback control system and convert the load related energy of the fluid motor and load means to an accumulated energy of the energy accumulating means for storing and subsequent use of this accumulated energy.
The high energy—efficient, load adaptive process of converting the load related energy to the accumulated energy is facilitated by regulating the exhaust fluid pressure drop across spool valve 2 by the energy recupturing pressure drop feedback control system and is basically controlled by the motor load. Note that the energy is being accumulated by the energy accumulating means, while the motor load is negative
( for example, during the deceleration of a motor vehicle ).
1 qad adaptive 5- The seconαVenergy converting means include the assisting supply line pressure drop feedback control system and convert the accumulated energy of the energy accumulating means to an assisting pressurized fluid stream being implemented for powering the supply power line L2 of spool valve 2. The assisting- pressurized fluid stream is actually gene'rated by an assisting variable delivery fluid power supply which is included into the assisting supply line pressure drop feedback control system and which is powered by the energy accumulating means. The high energy-efficient, load adaptive process of converting the nccu ulated energy to the assisting pressurized fluid stream is facilitated by regulating the assisting supply fluid pressure drop across spool valve 2 by the assisting supply line pressure drop feedback control system and is basically controlled by the motor load. Note that the energy is being released by the energy accumulating menns , while the motor load is positive ( for example, during the acceleration of the motor vehicle ).
6. Because the accumulation, storage, and release of the accumulated energy are all controlled by the motor load, the load adaptive energy regenerating system, as a whole, is also basically controlled by the motor load.
7. It can also be concluded that:
(a) the regeneration of a load related energy of the fluid motor and load means is facilitated by regulating the exhaust fluid pressure drop across valve 2 by the energy recupturing pressure drop feedback control system;
(b) the regeneration of a load related energy of the fluid motor and load means is also facilitated by regulating the assisting supply fluid pressure drop across valve 2 by the assisting supply line pressure drop feedback control system.
Regenerative adaptive fluid motor control system.
The above brief description of examplified regenerative adaptive fluid motor control systems (see iigures 11 to 22) can be still further generalized and extended by the comments as follows. 1. The primary supply line pressure drop feedback control system includes a primary variable delivery fluid power supply generating a primary pressurized fluid stream being implemented for powering the supply power line L2 of the spool valve 2. The assisting supply line pressure drop feedback control system includes an assisting variable delivery fluid power supply generating an assisting pressurized fluid stream being also implemented for powering the supply power line L2 of the sppo.l valve 2.
2. Note that assisting pressure rate P2 — PQ2 -(-ΔP^ of tne • assisting pressurized fluid stream is being correlated with the primary pressure rate 2 ==: Pr)2 T"ΔJP2 σf ^ne Pri'n ry pressurized fluid scream. Note also that P2R)ΔP2 and, tnerefore, P2R ^> 2 - wnile there is still any meaningful energy left in the energy accumulator.
3. In accordance with point 2, the assisting pressurized t lujd stream has a priority over the primary pressurized rluid stream in supplying the fluid power to the supply power line L2.
4. Speaking more generally, it can be concluded that regeneration of a load related energy of the fluid rnσtor and load means is accomadated by correlating the primary pressure rate of the primary pressurized fluid stream with the assisting pressure rate of the assisting pressurized fluid stream by regulating the primary supply fluid pressure drop across valve 2 and regulating the assisting supply fluid pressure drop across valve 2 by the primary supply line pressure drop feedback control system and the assisting supply line pressure drop feedback control system, respectively. ,
5. The exhaust line energy recupturing means of the energy recupturing pressure drop feedbyck control systems can be introduced by the exhaust line variable displacement motor
66 — see figures 11, 12, 16, 17, or by the exhaust line constant displacement motor 116 driving the exhaust line variable displacement pump 120 — see figures 18 to 22. 6. The assisting variable delivery fluid power supply, which is powered by the energy accumulating means, can be introduced by the ' assisting variable displacement pump 55 — see figures , 11, 12, 16, 1.7, or by. the assisting variable displacement motor 118 driving the assisting constant displacement pump 114 — see figures 18 to 21. The assisting variablle delivery fluid power supply can also be introduced by the assisting constant displacement motor 198 driving the assisting variable displacement pump 194 - as it is illustrated by Fig.22.
7. The primary variable delivery fluid power supply can be introduced by the primary variable displacement pump 90 - see figures 12, 19, and 22 or by the variable speed primary motor
( or engine ) 92 driving the primary fluid pump - see figures 11, 16, 17, 18, 20 and 21.
8. In accordance with points 5, 6, and 7 and the above description, any pressure drop regulation is accomplished by the related pressure drop feedback control system by implementing the related pressure drop feedback signal for modulating one of the following : a) the variable displacement means of the variable displacement pump , b) the variable displacement means of the variable displacement motor, c) the variable speed primary motor ( or the variable speed primary engine ) driving the primary fluid pump.
9. The variable displacement pumps having the built-in pressure drop feedback controllers are well known in the art. This type of control for the 'variable displacement pump is often called a "load sensing control" and is described in many patents and publica tioiros ( see, for example, Budzich — —U.S. Patent No , 07*., 529 of Feb.21, I978 ).
Moreover, the variable displacement pumps with the load-sensing pressure drop feedback controllers are produced (in mass amount ) by many companies which provide catalogs and other information on this load sensing control. Some of these companies are : a) THE OILGEAR COMPANY, 2300 South, 51th Street, Milwaukee, WI 3219, U.S.A. ( see, for example, Bulletin 47016A ); b) SAUER-SUNDSTRAND COMPANY, 2800 East 13th Street, Λmes IΛ 50010, U.S.A. ( see, for example, Bulletin 9825, Rev.E ) ,- c) DYNEX/ IVETT , INC., 770 Capitol Drive, Pewaukee,
WI 53072, U.S.A. ( see, for example, Bulletin PES-1289 ). Furthermore, the additional information of general nature on the feedback control systems and the hydraulic control systems is also readily available from many publications — see, for example, the books already named above. In short, the load-se-nsing pressure drop feedback controllers of the variable displacement pumps are, indeed, well known in the Art.
10. Comparing points -8 and 9 , it can be concluded that the load adaptive variable displacement means ( of the variable displacement pumps and the variable displacement motors ), which are used "in this invention, are basically similar with the well-known load-sensing pressure drop feedback controllers of the variable displacement pumps. These load adaptive displacement means can also be reffered to as the load adaptive displacement controllers.
11. It. is important to stress that load adaptive displacement means and the related pressure drop feedback' control systems, make it possible to eliminate the need for any special ( major ) energy commutating equipment.
Load adaptive displacement means and the energy regenerating circle.
Returning to Fig.22, let's consider more specifically the load adaptive displacement means 196 of pump 194 and the load adaptive displacement means 130 of pump 120. The examplified schematics of load adaptive displacement means 196 and 130 are presented by figures 23 and 24, respectively. These simplified schematics show:
(1) swashplates 246 and 266 of the variable displacement pumps; ϊ-
(2) swashplate hydraulic cylinders 242 and 262;
(3) forces Fg2 and Fgg of the swashplate precompressed springs;
(4) swashplate displacement restrictσrs 248 and 268;
(5) swashplate spool valves 250 and 270;
(6) the spool precompressed springs 254 and 274 defining command signals __ΔP R and j s. P , respectively; (7) the principal angular positions of swashplates ( "zero" angle, regulated angles, maximum angle, and small negative angle)
Figures 23 and 24 are simplified and made similar to the extend possible.
Reich swashplate is~driven"by a ϊunge ~bTl:. e related cylinder against the force of a precompressed spring. Each hydraulic cylinder is controlled by the related three-way spool valve which is also provided wj^th the pressure and tank lines.
The pressure line is powered by an input pressure P which is supplied by any appropriate pressure sourse.
The valve spool is driven by a pressure drop feedback signal against the force of the precompressed spring defining the pressure drop command signal.
Note that three-way valve can also be replaced by a two-way val e which does not have the tank line ( in "this case the tank line i connected through a throutle \ιal vs to the. line of hydraulic cylinder) .
The assisting supply line pressure drop feedback signal P2R~ f'()2 is applied to the spool 52 of valvθ 250 <see Fig.23) to construct the assisting supply line pressure drop feedback control system and, thereby, to maintain pressure P -= PQ^" _ P R in the outlet line 30 of the assisting constant displacement motor 198 which is driving the assisting variable displacement pump 194 (as it was already basically explained before). At the balance of the assisting supply line pressure drop feedback control, the spool 252 of valve 250 is in the neutral spool position which is shown on Fig.23. Note that •Δ. was already indicated before.
The exhaust line pressure drop feedback signal o5~p5 is applied to the spool 272 of valve 270 (see Fig.24) to construct the energy recupturing pressure drop feedback control system and. thereby, to maintain pressure Pr=Prιt.—ΔPC in the exhaust
5 05 5 line L5 powering the exhaust line constant displacement motor 116 which is driving the exhaust line variable displacement pump 120 (as it was already basically explained before).
At the balance of the exhaust line pressure drop feedback control. the spool 272 of valve 270 i The neutral spool position which is shown on Fig.24. Note that pressure drop command signals ^. P , ✓ . P2R> and ^Pc, are selected so that p ^.^ Z» < >) < as it is required by expression (3).
Fig.25 illustrates an examplified energy regenerating circle. It is assumed that the wheeled vehicle is moving in a horizontal direction only. As the vehicle is moving with a constant speed , decelerated, completely stoped, and accelerated, the related energy regenerating circle is completed. This stop-and-go energy regenerating circle has been already briefly introduced before ( to explain the concept of preventing a substantial pressure drop regulation interferrence ) and is easily readable on Fig. 5, when considered in conduction with figures 22 to 24 and the related text. For example, while the vehicle is decelerated, the swashplate 266 is positioned as indicated on Fig.24. While the vehicle is accelerated, the swashplate 246 is positioned as indicated on Fig.23.
Regenerative drive system having the combined energy accumulating means.
The schematic shown on Fig..19 is now further modified to replace the independent regenerating circuitry by the built-in regenerating circuitry and to improve the utilization of the combined energy accumulating means. Accordingly, the assisting variable delivery fluid power supply (motor 118 driving pump 114), the check valves 40 and 44, and the electrohydraulic energy converting means 142 are eliminated. The modified schematic is shown on Fig.26. The added components are:
(a) electrical motor-generator 290, (b) constant displacement motor 300, (c) shut-off valve 298, and (d) check valve 296. The primary engine (motor) 100, the direct-current motor-generator 290, the hydraulic motor 300, and the hydraulic pump 90 are all mechanically connected by a common shaft 98. The motor-generator 290 is also electrically connected (through lines 292 and 294) with the electrical accumulator 144. On the other hand, the hydraulic accumulator 122 is hydraulically connected (through shut-off valve 298) with the inlet line 302 of motor 300.
The regenerative drive system of Fig.26 makes it possible to minimize the required engine size of a wheeled vehicle. The engine 100 is provided with a speed control system which is assumed to be included in block 100 and which is used to maintain a preselected (basic) speed of shaft 98 while allowing some speed fluctuations under the load which is applied to the shaft 98. The related margin of accuracy of the speed control system is actually used to maitain a balance of power on the common shaft 98 and, thereby, to minimize the required engine size of a wheeled vehicle.
The driving torque of shaft 98 is generally produced by engine 100, by motor-generator 290 (when it is working as a motor), and by motor 300 (when it is powered by the hydraulic accumulator 122 through shut-off valve 298). The loading torque of shaft 98 is basically provided by pump 90 and by motor-generator 290 (when it is working as a generator). Note that at some matching speed of shaft 98 (within the margin of accuracy of the speed control system) a speed-dependent voltage of generator 290 is equal to a charge-dependent voltage of accumulator 144, so that no energy is transmitted via lines 292 and 294. As the speed of shaft 98 is slightly reduced, the electrical energy is transmitted from the electrical accumulator 144 to the electrical motor 290 helping engine 100 to overcome the load. On the other hand, as the speed of shaft 98 is slightly increased, the electrical energy is transmitted from the electrical generator 290 to the electrical accumulator 144 , allowing to utilize the excess power of shaft 98 for recharging the electrical accumulator 144. Note also that shut-off valve 298 is normally closed and is open only under some preconditions - in order to power the constant displacement motor 300 by the hydraulic energy of accumulator 122. Let's assume, first, that a wheeled vehicle, such as a city transit bus, is moving in a horizontal direction only. And let's consider briefly the related energy regenerating circle. bus
1. Λs thesis moving with a constant speed, the pump 90 is basically powered by engine 100.
2. Λs the bus is decelerated, the mechanical energy of a bus mass is converted to the hydraulic energy of accumulator 122. The excess energy of accumulator 122 is converted - via valve 298, motor 300, and generator 290 - to the electrical energy of accumulator 14 . The primary engine 100 may also participate in recharging the electrical accumulator 144.
3. Λs the bus is stoped, the engine 100 is used only for recharging the electrical accumulator 144.
4. Λs the bus is accelerated, the pump 90 is basically powered by motor 300 and is also powered by engine 100 and motor 290. The constant displacement motor 300 is powered by the hydraulic accumulator- 122, through shut-off valve 298.
Λs the bus is moving up-hill, the pump 90 is driven by engine 100 and motor 290 which is powered by electrical accumulator 144. Λs the bus is moving down-hill, the mechanical energy of a bus mass is converted to the hydraulic energy of accumulator 122, and this hydraulic energy is further converted - via valve 298, motor 300, and generator 290 - to the electrical energy of accumulator 144.
Λn optional control signal "S" which is applied to the shut-off valve 298, is produced by an optional control unit which is not shown on Fig. 6. This control unit can be used for controlling such optional functions as follows:
(a) opening shut-off valve 298 - when the vehicle is accelerated;
(b) opening shut-off valve 298 - when the vehicle is moving down-hill and after accumulator 122 is substantially charged;
(c) opening shut-off valve 298 - when the vehicle is decelerated, in order to convert the excess energy of hydraulic accumulator 122 to the electrical energy of accumulator 144; (d) opening shut-off vn.lve 290 juθt alter nccυmυ.lotor 122 J.π w..ιl .-.nnl..ln!...y charged .
It. πhoulcl be emphasi e that schematic of Fig.20 J.Θ of o very enorβ.l. nature. The <_jκ ιnp.l.i.C.l9d mo ifications of this schema ic een be briefly cl.orαoUeri θd am follows:
(a) the constant displacement motor 300 is of a smaller flow capacity in comparison with the variable displacement pump 90;
(b) the variable displacement pump 90 is also used as a motor to provide an al ternative route for transmission of energy from accumulator 122 to the common shaft 98;
(c) providing at least two preselectable (basic) speeds of shaft 98 to respond to the changing load invironments;
(d) modifying the hybrid motor means driving pump 90 — as it is explained at the end bf this description.
Two basic types of regenerative systems.
There are basically two types of regenerative adaptive fluid motor control systems: (a) the regenerative system having an independent regenerating circuitry (see figures 11 to 22) and (b) the regenerative system having a built-in energy regenerating circuitry (see figures 9, 10. and 26). The first type of regenerative systems is identified by that the primary and assisting supply line pressure drop feedback control systems are separated. The second type of regenerative systems is identified by that the primary and assisting supply line pressure drop feedback control systems are not separated and are represented by only one supply line pressure drop feedback control system. The generalized first-type systems have been already introduced by figures 13, 14, and 15. Λ generalized second- ype system is shown on Fig.27, which is mostly self-explanatory and is still further understood when compalred with figures 9, 10, 26, and 15. Note that transition from the first to the second type of regenerative systems is accomplished typically by replacing the separated primary and assisting supply line pressure drop feedback control systems by only one supply line pressure drop feedback control system and by implementing the primary power supply means for powering the energy accumulating means. For example, in the regenerative system of Fig.22, the transition from the independent regenerating circuitry to the built-in regenerating circuitry can be accomplished by eliminating the separated primary supply line pressure drop feedback control system and by implementing the primary pump 90 for powering the hydraulic accumulator 122 ( the resulted schematic can be still further modified to incorporate also an electrical accumulator ) .
The two basic types of regenerative systems can generally be combined to include both - the built-in regenerating circuitry and the independent regenerating circuitry. For example, in the regenerative system of Fig.26, the transition to the combined schematic can be accomplished by adding an assisting supply line pressure drop feedback control system, which is shown on Fig.22 and which includes the constant displacement motor 198 driving the variable displacement pump 194. The resulted combined schematic is also applicable to the wheeled vehicles .
Adaptive lluld control and the load n Lhonmenlg
|t Is understood that thin Invention la not . linjled to any μ.ιι Ucul-u- a lica i n. It o to way that, figures 1. 4. 9 . and \ 2 , are not relate onl t the hydraulic pr mises. II IM Λlm. to say that J lyurββ 10. .1 J . .16 to 22. and 26 o e IUJI rof.il. l only to the motor vel.lc.len. In fact, the typical ad,=u>l J ve schematics which are shown on figures .1 , 4, 6, 9 to 12, 16 to 22, and 26 are exluslvely associated only with a type of motor load of the fluid motor 1 (or 150), as it is characterized below:
(a) the shematics shown on figures 1, 4,' 9, and 12 are adatJve to the one -directional static load force:
(b) the schematics shown on figures 10, ll, 15 to 22, and 26 are adaptive to the two-directional dynamic load force, which is generated during acceleration and deceleration of a load mass moving only in one direction..
(Q) thβ schematic shown on Pig.6,ia adaptive to thβ two-directional static load forces .
simplified Thβ above .load-related classification of typical adaptive schematics is instrumental in modifying these schematics for tho modified load environments.
For example, the schematic shown. on Fig.18 is adaptive tc the two-directional dynamic load force, which is generated during acceleration and deceleration of a load mass moving only in one direction. If the load environments are modified by replacing this two-directional dynamic load force by the one-directional static lo«ιd force, the schematic of Fig.15 must be also modified. The modified schematic may include he five-way spool valve 2 instead of the four-way spool valve 2 which is shown on Fig.18. In this case, the energy regenerating circuitry using hydraulic accumulator 122 must be switched over from the exhaust o e line L5 to the exhaust power line L . as it is illus-ratr by Fig.12 — for a case of using the flywheel accumulat :- Q .
Adaptive fluid control with a supplementary output motor.
There is one special modification of independent regenerating circuitry which is not covered by the generalized schematic of Fig.15 and which, therefore, is considered below. The regenerative braking pump 170 of Fig.21 can also be used as a variable displacement motor to make-up a supplementary variable displacement motor/pump. The pump functions of this supplementary output motor/pump have been already studied with the help of Fig.21. For simplicity, the motor functions of this supplementary output motor/pump will also be studied - separately.
Fig.28 is derived from Fig.21 by replacing the supplementary output pump 170 by the supplementary output motor 170 and by eliminating the assisting supply line pressure drop feedback control system (including motor 114 and pump 118 ) and some other components (check valves 40, 44, and 174). The variable displacement motor 170 is powered by the hydraulic accumulator 122 through a shut-off valve 297 which is basically controlled • by pressure signal Pg2. While this pressure signal is comparatively small, the shut-off valve 297 is closed. Λs signal PQ2 IS further raising-up, the shut-off valve 297 is open, provided that there is still enough energy stored in the hydraulic accumulator 122.
The variable displacement means 99 of motor 170 are constructed to make-up a displacement feedback ■ control system including a variable displacement mechanism (of motor 170) a displacement feedback control errow signal Δd, generated in accordance with a difference between
3 command-displacement signal do = C » Pn0~2 ( where Cp is a constant coefficient ) and a mechanism displacement ( feedback signal ) d. of the variable displacement mechanism of motor 170, Λ pressure-displacement transducer converting the pressure signal PQ2 into the proportional command-displacement signal d Q? is included into the variable displacement means 99 of motor 170. Tli is transducer may incorporate, for example, a small spring-loaded hydraulic cylinder actuated by the pressure signal p
02- The displacemen feedback control errow signal ,A.d = d — d, is implemented for modulating the variable displacement mechanism of motor 170 for regulating the mechanism displacement d-. of the variable displacement motor mechanism of ""▼ 170 in accordance with the command sigbnal do
( and hence, in accordance with the pressure rate 02 _== d0/CP„ in the motor line Ll ).
It should be emphasized that the displacement feedback control system, which is well known- in the art, is, in fact, the position feedback control system and that, therefore, the general position feedback control technique, which is characterised above with respect to the fluid motor position feedback control system, is also basically applicable to the displacement feedback control system.
In general, the displacement feedback control circuitry of motor 170 is adjusted so that, while the pressure PQ2 in the motor line Ll is comparatively low, this circuitry is not operative and d^=0. As the pressure P* in the
02 motor line Ll is further ralsing-up, the displacement d of motor 170 is increasing acdordingly, so that the total accelerating torque is properly distributed between the fluid motor 1 and the supplementary motor 170. The use of motor 170 makes it possible to substantially increase the available (total) accelerating torque of the wheeled vehicle.
Note that the use of motor 170 on small displacements should be avoided for as much as possible. Note also that a significant dynamic performance superiority must be provided for the displacement feedback control system against the primary supply line pressure drop feedback control system, in order to prevent their substantial dynamic operation interference. The concept of providing "a significant dynamic performance superiority" have been already generally introduced before and is further readily applicable to the displacement feedback control system versus the primary supply line pressure drop feedback control system.
The related generalized schematic of Fig.29 is derived from Fig.15. is mostly self-explanatory, and is reflective of the facts that the assisting supply line pressure drop feedback control system is now eliminated and that pressure Pac from the hydraulic accumulator 122 is now applied to the supplementary output motor 170 of the fluid motor and load means.
Some other related considerations.
The schematic of Fig.26 can be modified by changing the hybrid motor means driving pump 90. The examplified modifications are as follows.
1. The electrical motor-generator 290 and the related electrical accumulator 144 are exluded from this schematic. , The constant displacement motor 300 is replaced by a variable displacement motor which is used to construct a supplementary shaft-speed feedback control system maintaining the preinstalled speed of shaft 98 when this variable displacement motor is powered by accumulator 122. As a result, the hydraulic energy of accumulator 122 is transmitted to shaft 98 in accordance with the actual energy requiren ent . Note that possible interference between the main shaft-speed feedback control system ( of primary engine 100 ) and the supplementary shaft-speed feedback control system ( of the variable displacement motor ) is prevented by providing
V =_= V -- ΛV , CS CM -*• ;
where:
VCM — is a velocity command-signal for the main shaft-speed feedback control system,
Vcs — is a velocity command—signal for the supplementary shaft-speed feedback control system, and
ΔV — is a sufficient velocity margin between these two systems.
In other words, the supplementary speed control system should actually be regulated just "slightly above" the main speed control system.
2. The primary engine 100 is excluded from the schematic of Fig.26. In this case, the primary energy should be supplied by the electrical accumulator 144.
3. The primary engine 100 is disconnected from shaft 98 and is driving a constant displacement pump which is powering the constant displacement motor 300. In this case, the hydraulic energy of accumulator 122 is transmitted to shaft 98 via this constant displacement pump driving the constant displacement motor 300.
The schematic of Fig.22 can be modified by providing the primary engine 100 with a variable-speed feedback control system which is used for maintaining the engine maximum energy efficiency. Note that as the engine speed increases, the displacement of pump 90 is being reduced accordingly, to maintain the pump flow output which is defined only by the opening of valve 2.
The schematic of Fig.22 can also be modified by eliminating the primary supply line pressure drop feedback control system
( like it is ) and by implementing the primary pump 90 for powering the hydraulic accumulator 122. The resulted schematic having a built-in energy regenerating circuitry can also be constructed for maintaining the engine maximum energy efficiency.
It should be emphasized that adaptive fluid control schematics being considered are the concept illustrating schematics only. Some design related considerations are as follows.
1. The maximum and minimum pressures in any fluid power line must be restricted,.
2. The primary supply power line 54 (see figures .1.1 to 22) can be protected by the maximum pressure relief valve.
The maximum pressure in line 54 can also be restricted by using the variable delivery means 93 of pump 90.
In general, the maximum pressure relief valves can also be used to protect other hydraulic lines.
3. The check valve 154 (figures 20 and 22) is added to very efficiently restrict the maximum pressure in the exhaust motor line 1,4 by relieving an excess fluid from this line (through check valve 154) into the high-pressure hydraulic accumulator 122.
4. Similar to point 3, the check valves can be used to restrict the maximum pressure in still other power lines.
5. The check valve 155 (figures 20 and 22) is added to effectively restrict the minimum pressure in the supply motor line hi by connecting this line (.through check valve 155) with the tank 62. ' .
6. Similar to point 5, the check valves can be used to restric the minimum pressure in still other power lines.
For example, the exhaust power line 1.5 (or L3) should usually be connected through a check valve to the tank to avoid creating a vacuum in this line.
7. The oil tank capacity can often be reduced, the oil cooling system can often be eliminated.
8. The oil tank 62 can often be replaced by a low-pressure hydraulic accumulator (accompanied by a small-supplementary tank ) .
9. The oil tank 62 can also be supplemented by a low-pressure centrifugal pump.
How to restrict a supply line power rate.
Still other engineering consideration on a way of transition from the concept illustrating schematics to the practical design of regenerative adaptive fluid motor control systems is how to restrict the supply line power rate — when it is needed. Let's consider, for example, the schematic of Fig.26 — as it is applied to the motor vehicles. In this case, the required supply line power rate in line L2 is defined by the load 96 of motor 1 and by the opening of valve 2 and can generally exceed the combined power supply capacity of engine 100 and electrical motor 290. To prevent this overload event from happening, the practical design must include the means of restricting the spool displacement (SD) of valve 2 versus the load pressure rate (LP) in line Ll ( or in line L2 ), so that the resulted load power rate ( which is proportional to LP x SD ) would not exceed the limited power supply capacity.
In other words, the practical regenerative adaptive fluid motor control systems may include the means of restricting the required load power rate in accordance with the limited power supply capacity.
Regenerative adaptive fluid control versus Non-regenerative adaptive fluid control.
As it was already mentioned before, there are basically two types of the two-way load adaptive fluid motor control systems . The non-regenerative adaptive fluid motor control systems are equipped with an exhaust line pressure drop feedback control system including an exhaust line pressure drop regulator. On the other hand, the regenerative adaptive fluid motor control systems are equipped with an energy recuperating pressure drop feedback control system including an exhaust line energy recuperating means .
The above description is presented in a way of transition from the non-regenerative adaptive fluid motor control to the regenerative adaptive fluid motor control . Note that the resulted regenerative adaptive fluid control schematics being considered are, in fact, convertable . The transition from these schematics back to the non-regenerative adaptive fluid control can generally be accomplished by replacing the energy recuperating pressure drop feedback control system ( and the related energy regenerating circuitry ) by the • exhaust line pressure drop feedback control system including the exhaust line pressure drop regulator.
While my above description contains many specificities, those should not be construed as limitations on the scope of the invention, but rather a.s an exemplification of some preferred embodiments thereof. Many other variations are possible. For example, the schematic shown on Fig can be easily modified to convert the five-way valve 2 to the six-way valve by separating the supply power line L6 from the supply power line L2. The separated line L6 can be then connected directly to the line 5^ of the additional hydraulic power supply 5° shown on Fig. .
Various modifications and variations, which basically rely on the teachings through which this disclosure has advanced the art, are properly considered witnin the scope of this invention as defined by the appended claims and their legal equivalents .

Claims

Claims to
"Regenerative adaptive fluid control'.'
What is claimed is:
1. Λ regenerptive adaptive fluid motor control method comprising the steps of :
constructing a fluid motor control system including fluid motor and load means, spool valve means, and fluid power means; said fluid motor and load means including fluid motor means and motor load means and accumulating a load related energy; said spool valve means having at least three fluid power lineq including a motor line conducting a motor fluid i o to or a motor fluid flow from said fluid motor means of said fluid motor and load means, a supply power line, and an exhaust power line; said fluid power means including a primary variable delivery fluid power supply generating a primary pressurized fluid stream being implemented for powering said supply power line of said spool valve means;
introducing an exhaust line energy recuperating means for varying a counterpressure rate in said exhaust power line and for recuperating said load related energy of said fluid motor and load means; constructing an energy recuperating pressure drop feedback control system including said exhaust line energy recuperating means; regulating an exhaust fluid pressure drop across said spool valve means by said energy recuperating pressure drop feedback control system by varying said counterpressure rate in said exhaust power line by said exhaust line energy recuperating means;
constructing a load adaptive energy regenerative), system including first load adaptive energy converting means, energy accumulating means, and second load adaptive energy converting means; said first load adaptive energy converting means including said energy recuperating pressure drop feedback control system;
providing a load adaptive regeneration of said load related energy of said fluid motor and load means by said load adaptive energy regenerating system by converting said load related energy through said first load adaptive energy converting means including said energy recuperating pressure drop feedback control system to a recuperated energy of said energy accumulating means, by storing said recuperated energy by said energy accumulating means, and by converting said recuperated energy through said second load adaptive energy converting means to a regenerated energy reusable by said fluid motor and load means; facilitating said load-adaptive regeneration of said load related energy of sai.d fluid motor and load means by regulating said exhaust fluid pressure drop across said spool valve means by said energy recuperating pressure drop feedback control system
constructing a primary supply line pressure drop feedback control system including said primary variable delivery fluid power supply; regulating a primary supply fluid pressure drop across said spool valve means by said primary supply line pressure drop feeback control system by varying a primary pressure rate of said primary pressurized fluid stream by a primary variable delivery means of said primary variable delivery fluid power supply.
2. The method according to claim l , wherein said exhaust line energy recuperating means includes ^π exhaust line variable displacement motor being powered by s id exliaust power line,
HIid wherein varying said counterpressure rate in said exliaust power line is accomplished by an exhaust line variable displacement means of said exhaust line variable displacemenl: in o t o r .
3- 'The method according to claim 1 wherein said exhaust line energy recuperating means includes *" exhaust line fluid motor ein powereci by said exhaust power line and driving •-.ii exhaust line variable displacement pump, and wherein varying said counterpressure rate i„ said exhaust power line is accomplished by an exhaust line variable displacement means of said exhaust line variable d splacement pump.
4.
The method according to claim 1 wherein said fluid motor means include at least one hydraulic cylinder having at least one loadable chamber, and wherein said load related energy of said fluid motor and load means includes a compressed fluid energy of said loadable chamber of said hydraulic cylinder.
5. The method according to claim 1 wherein' said motor load means include a frame of a hydraulic press, wherein said fluid motor means include at least one hydraulic cylinder of said hydraulic press, wherein said hydraulic cylinder includes a loadable chamber being loaded against said frame of said hydraulic press, and wherein said load related energy of said fluid motor and load means includes a compressed fluid energy of said loadable chamber of said hydraulic cylinder of said hydraulic press.
The method according to claim 1 wherein said motor load means include a mass load of said fluid motor means, and wherein said load related energy of said fluid motor and load means includes a mechanical energy of a mass of said mass load.
7'• The method -according to claim 1 wherein said motor load means include a mass of a wheeled vehicle, wherein said fluid motor means are loaded by said mass of said wheeled vehicle, and wherein said load related energy of said fluid motor and load means includes a mechanical energy of said mass of said wheeled vehicle. 8- The method according to claim wherein said primary variable delivery fluid power supply includes a primary variable displacement pump generating said primary pressurised' fluid stream, • and wherein varying said primary pressure rate of said primary pressurized fluid stream is accomplished by a primary variable displacement means of said primary variable displacement pump.
9. The method according to claim 1 wherein said primary variable delivery fluid power supply includes a primary variable speed motor driving a primary fluid pump generating said primary pressurized fluid stream, and wherein varying said primary pressure rate of said primary pressurized fluid stream is accomplished by said primary variable speed motor.
10. The method according to claim 1 wherein said energy accumulating means are implemented for powering an assisting variable delivery fluid power supply generating an assisting pressurized fluid stream being implemented for powering said fluid motor means through said spool valve means,- wherein said second load adaptive energy converting means include an assisting supply line pressure drop feedback control system containing said assisting variable delivery fluid power supply and regulating an assisting supply fluid pressure drop across said spool valve means by varying an assisting pressure rate of said assisting pressurized fluid stream by an assisting variable delivery means of said assisting variable delivery tluid power supply, and wherein said method further comprising : accommodating said lpad adaptive regeneration of said load related energy of said fluid motor and load means by correlating said primary pressure rate of said primary pressurized fluid stream with said assisting pressure rate of said assisting pressurized flurd stream by regulating said primary supply fluid pressure drop across said spool valve means and regulating said assisting supply fluid pressure drop across said spool valve means by said primary supply line pressure drop feedback control system and said assisting supply line
pressure drop feedback control. system, respectively. 11 • The method according to claim 10 wherein said energy accumulating means include a flywheel, wherein said assisting variable delivery fluid power supply includes an assisting variable displacement pump being driven by said flywheel and generating said assisting pressurized fluid stream, and wherein varying said assisting pressure rate of said assisting pressurized fluid stream is accomplished by an assisting variable displacement means of said assisting variable displacement pump.
12. The method 'according to claim ιo wherein said energy accumulating means include a hydraulic accumulator, wherein said assisting variable delivery fluid power supply includes an assisting fluid motor being powered by said hydraulic accumulator and driving an assisting variable displacement pump, wherein said assisting pressurized fluid stream is represented by an exhaust from said assisting fluid motor, and wherein varying said assisting pressure .rate of said assisting pressurized fluid • stream is accomplished by an assisting variable displacement means of said assisting ariable displacement pump. i3- The method according to claim 10 wherein said energy accumulating means include a hydraulic accumulator, wherein said assisting variable delivery fluid power supply includes an assisting variable displacement motor being powered by said hydraulic accumulator and driving an assisting fluid pump generating said assisting pressurized fluid stream, and wherein varying said assisting pressure rate of said assisting pressurized fluid stream is accomplished by an assisting variable displacement means of said assisting variable displacement motor.
14. The method according to claim 1 wherein said fluid motor means include a variable displacement motor, and wherein said method further comprising: constructing a displacement feedback control system including a variable displacement mechanism of said variable displacement motor; regulating a mechanism displacement of said variable displacement mechanism of said variable displacement motor by said displacement feedback control system at least approximately in accordance with a mechanism displacement command signal being correlated with a spool displacement signal of said spool valve means.
15. The method according to claim 1 further comprising i constructing a, fluid' motor output feedback control system including said fluid motor control system and having output feedback control means measuring a motor output υf said fluid motor means and producing an output feedback control error signal regulating said motor output of said fluid motor means by said fluid motor output feedback control system by modul?ting said spool valve means by said output feedback control error signal ι preventing a subs antial..' dynamic operation i ni.erfei-pπce between •' regulating said exliaust fluid pleasure drop and regulating said motor output by providing a significant, dynamic performance superiority for said energy recuperating pressure drop feedback control system against said fluid motor output : feedback control system by providing either a significant frequency-response superiority or a significant final-transient-time superiority for said energy recuperating pressure drop feedback control system against said fluid motor output feedback control system;
supply line~p essure dr p feedback control system against sa fluid motor output feedback .control system.
The method according to claim 15
~,n ssid n-« -t-r output feedback -trol s ste »
. A. - ,-_,__,_-_ hnck control n ed by a fluid motor position feedbaci represei s ys em . onted ά output .feedback' control means are repres where ii: n sa. by posi .t i i ■on n cnonπv.rioul means, wherein said .otor output is represented by a -tor
herein said output feedback control error signal is a epres ented by a position feedback control error signal.
^7- A regenerative adaptive fluid motor control system comprising:
a fluid motor control system including fluid motor and load means, spool valve means, and fluid power means . said fluid motor and load means including fluid motor means and motor load means and ..accumulating a load related energy; said spool valve means having, at least three fluid power line'? including a motor' line conducting a motor fluid flow to or a motor fluid flow from .said fluid motor means of sal fluid motor and load means, a supply power line, and an exhaust powe line; said fluid power means inclμding a primary variable delivery fluid power supply generating a primary pressurized fluid stream being implemented for powering said supply power line of said spool valve means j an exhaust line energy recuperating means for varying a counterpressure rate in said exhaust power line and for recuperating said load related energy of said fluid motor and load means; an energy recuperating pressure drop feedback control system including said exhaust line energy recuperating means and operable to regulate an exhaust fluid pressure drop across said spool valve means by varying said counterpressure rate in said exhaust power line by said exhaust line energy recuperating means; a load adaptive energy regenerating system including first load adaptive energy converting means, energy accumulating means, and second load adaptive energy converting means; said first load adaptive energy converting means including said energy recuperating pressure drop feedback control system and operable to convert said load related energy of said fluid motor and load means to a recuperated energy of said energy accumulating means for storing and subsequent use of said recuperated energy; said second load adaptive energy converting means operable to convert said recuperated energy of said energy accumulating means to a regenerated energy reusable by said fluid motor and load means;
a primary sup'ply line pressure drop feedback control system including said primary variable
.operable to delivery fluid power supply and v egula e a primary supply fluid pressure drop across said spool valve means by varying a primary pressure rate of said primary pressurized fluid streaπi by a primary variable delivery means of said primary variable delivery fluid power supply.
18. The system according to claim 17 wherein said energy accumulating means are implemented for powering an assisting variable' delivery fluid power supply generating an assisting pressurized fluid stream being implemented for powering said fluid motor means through said spool valve means, and wherein said second load adaptive energy converting means include an assisting supply line pressure drop feedback control system containing said assisting variabJe delivery fluid power- supply and operable to regulate an assisting supply fluid pressure drop across said spool valve means .by varying an assisting pressure rate of said assisting pressurized fluid stream by an assisting variable delivery means of said assisting variable delivery tluid power supply^
19. Λ regenerptive pdnptive fluid motor control method comprising the steps of .
constructin a fluid motor control system including fluid motor and load means, spool valve means, and fluid power means ; said fluid motor and load" means including fluid motor means and motor load means and accumulating a load related energy; said spool valve means having at least three fluid power line*? including a motor line conducting a motor fluid flow to or a motor fluid flow from said fluid motor means of said fluid motor and load means, a supply power line, and an exhaust power line; said fluid power means including energy accumulating means being implemented for powering a variable delivery fluid power supply generating a pressurized fluid stream being implemented for powering said supply power line of said spool valve means; 1 "
introducing an exhaust line energy recuperating mearF for varying a counterpressure rate-in said exhaust power line and for recuperating said load related energy of said r.lu.id motor and load means; constructing an energy recuperating pressure drop feedback control system including said exhaust line energy recuperating means; regulating an exhaust fluid pressure drop across said spool valve means by said energy recuperating pressure drop feedback control system by varying said counterpressure rate in said exhaust power line by said exhaust line energy recuperating means;
constructing a supply line pressure drop feedback control system including said variable delivery fluid power supply; regulating a supply fluid pressure drop across said spool valve means by said supply line pressure drop feedback control system by varying a pressure rate o said pressurized fluid stream by a variable delivery means of said variable delivery fluid power supply; constructing a load adaptive energy regeneratiλ^ system including first load adaptive energy converting meahs, said energy accumulating means, and second load adaptive energy converting means; said first Joad adaptive energy converting means including said energy recuperating pressure drop feedback control 'system; said second load adaptive energy converting moans including said supply line pressure drop feedback control system; providing a load adaptive regeneration of said load related energy of said fluid motor and load means by said load adaptive energy regenerating system by converting said load related energy through said first load adaptive energy converting means including said energy recuperating pressure drop feedback control system to a recuperated energy of said energy accumulating means, by storing said recuperated energy by said energy accumulating means, and by converting said recuperated energy through said second load adaptive energy converting means including said supply line pressure drop feedback control system to a regenerated energy of said pressurized fluid stream being implemented for powering said supply power line of said spool valve means; facilitating said load adaptive regeneration of said load related energy of said fluid motor and load means by regulating said exhaust fluid pressure drop across said spool valve means and regulating said supply fluid pressure drop across said spool valve means by said energy recuperating pressure drop feedback control system and said supply line pressure drop feedback control system, respectively.
20. The method according to claim 19, wherein said fluid power means include a primary power supply being implemented for powering said variable delivery fluid power supply. and wherein a primary energy of said pressurized fluid stream is supplied by said primary power supply through said variable delivery fluid power supply.
21- The method according to claim 19 , wherein said fluid power means include a primary power supply being implemented for powering said energy accumulating means , and wherein a primary energy of said pressuried fluid stream is supplied by said primary power supply through said energy accumulating means.
22. T e method according to claim 19 , wherein said exhaust line energy recuperating means includes an exhaust line variable displacement motor being powered by ηjd exliaust power line, n wherein varying said counterpressure rate in said exhaust power line is accomplished by an exhaust line variable displacement means of said exhaust line variable displacement in o tor.
23. -The method according to claim 19 wherein said exhaust line energy recuperating means includes 3iι exhaust line fluid motor being powered by said exhaust power line and driving .-'ii exhaust line variable displacement pump, and therein varying said counterpressure rate in said exhaust power line is accomplished by -an exhaus line variable displacement means of. said exhaust line variable dis lacement pump.
4. The method according to claim 19 wherein said fluid motor means include at least one hydraulic cylinder having at least one loadable chamber, and wherein said load related energy of said fluid motor and load means includes a compressed fluid energy of said loadable chamber of said hydraulic cylinder.
25. The method according to claim 19 « wherein said motor load means include a frame of a hydraulic press, wherein said fluid motor means include at least one hydraulic cylinder of said hydraulic press, wherein said hydraulic cylinder includes a loadable chamber being loaded against said frame of said hydraulic press, and wherein said load related energy of said fluid motor and load means includes a compressed fluid energy of said loadable chamber of said hydraulic cylinder of said hydraulic press.
26. The method according to claim 19 , wherein said motor load means include a mass load of said fluid motor means, and wherein said load related energy of said fluid motor and load means includes a mechanical energy of a mass of said mass load.
2V. The method ^according to claim 19 , wherein said motor load means include a mass of a wheeled vehicle, wherein said fluid motor means are loaded by said mass of said wheeled vehicle, and wherein said load related energy of said fluid motor and' load means includes a mechanical energy of said mass of said wheeled vehicle. 28, The method according to claim 19 , wherein said energy accumulating means include a flywheel, wherein said variable delivery fluid power supply includes a variable displacement pump being driven by said flywheel and generating. said pressurized fluid stream, and wherein varying said pressure rate of said pressurized fluid stream is accomplished by a variable displacement means of said variable displacement pump.
29 The method according to claim 19 , wherein said energy accumulating means include a hydraulic accumulator. wherein said variable delivery fluid power supply includes a fluid motor being powered by said hydraulic accumulator and driving a variable displacement pump, . wherein said pressurized fluid stream is represented by an exhaust from said fluid motor, and wherein varying said pressure rate of said pressurized fluid stream is accomplished by a variable displacement means of said variable displacement pump. 30 • The method according to claim 19 . wherein said energy accumulating means include a hydraulic accumulator , wherein said • variable delivery fluid power supply includes a variable displacement motor being powered by said hydraulic accumulator and driving a fluid pump generating said pressurized fluid stream, and wherein varying said pressure rate of said pressurized fluid stream is accomplished by a. variable displacement means of said variable displacement motor.
116
31. The method according to claim 19 wherein said variable delivery fluid power supply includes a variable displacement pump generating said pressurized fluid stream, wherein said fluid power means include a primary motor being implemented for driving said variable displacement pump, wherein said energy accumulating means are implemented for powering said variable displacement pump, and wherein varying said pressure rate of said pressurized fluid stream is accomplished by a variable displacement means of said variable displacement pump.
32. The method according to claim 19 wherein said variable delivery fluid power supply includes a variable displacement pump generating said pressurized fluid stream, wherein said variable displacement pump is driven by hybrid motor means including an energy regenerating fluid motor, wherein said energy accumulating means include fluid energy accumulating means being impilemented for powering said energy regenerating fluid motor,
and wherein varying said pressure rate of said pressurized fluid stream is accomplished by a variable displacement means of said variable displacement pump. 33- The method according to claim 19 wherein said variable delivery fluid power supply includes a variable displacement pump generating said pressurized fluid stream, wherein said variable displacement pump is driven by an electrical motor, wherein said energy accumulating means include fluid energy accumulating means being implemented for powering electrical energy accumulating means being implemented for powering said electrical motor, and wherein varying said pressure rate of said pressurized fluid stream is accomplished by a variable displacement means of said variable displacement pump.
34. The method according to claim 19 wherein said fluid motor means include a variable displacement motor, and wherein said method further comprising: constructing a displacement feedback control system including a variable displacement mechanism of said variable displacement motor; regulating a mechanism displacement of said variable displacement mechanism of said variable displacement motor by said displacement feedback control system at least approximately in accordance with a mechanism displacement command signal being correlated with a spool displacement signal of said spool valve means.
35. The method according to claim 19 further comprising . quoting a fluid motor output feedback control cons' tern including said fluid motor control system and having sys tput feedback control means measuring a motor output ou t.r ,„!- fluid «tor mean and Producing an output feedback
control error signal*, relating said motor output of said fluid motor means by said fluid motor output feedback control system by „odul,ting said spool valve means by said output feedback
control error signal : preventing a subs tantiέ i.' dynamic operation i. n 1: e rr.fereπce between : regulating said exhaust fluid pressure drop and regulating said motor output by providing a significant, dynamic performance superiority rot- said energy recuperating pressure drop feedback control system against said fluid motor output '. feedback control system by providing either a significant frequency-response superiority or a significant final-transient-time superiority for said energy recuperating pressure drop feedback control system against said fluid motor output feedback control system; a significant final-transient-time superiority for said supply line pressure drop feedback control system against said fluid motor output feedback .control system.
36 The method according to claim 35 .
!ln said fluid motor output feedback control system is wh ere :
,. , bv a fluid motor position feedback control represented oy a I uu system, wh6rein said output .feedback' control means are represented b »yy position feedback control means, herein said motor output is represented by a motor
osition, and wherein said output feedback control error signal is represented by a position feedback control error signal-
37. Λ regenerative nd»ptive fluid motor control system comprising ...
a fluid motor control system including fluid motor and load means, spool valve means, and fluid power means. said fluid motor and load means including fluid motor means a d motor load means and accumulating a load related energy. said spool valve means having at least three fluid power lineq including a motor line conducting a motor fluid flow to or a motor fluid flow from said fluid motor means of said fluid motor and load means, a supply power line, and an exhaust power line; said fluid power means including energy accumulating means being implemented for powering a variable delivery fluid power supply generating a pressurized fluid stream being implemented for powering said βupply power line of said spool valve means; an exhaust line energy recuperating means for varying a counterpressure rate in said exhaust power 1 ine and for recuperating said load related energy of said fluid motor and load means; an energy recuperating pressure drop feedback control system including said exhaust line energy recuper ting means and operable to regulate an exhaust fluid pressure cjrop across said spool valve means by varying said counterpressure rate in said exhaust power line by said exhaust line energy recuperating means; a, supply line pressure dr<|>p feedback control system including said; variable delivery fluid power supply and operable to regulate a supply fluid pressure drop across said spool valve mean by varying a pressure rate of said pressurized fluid stream by a variable delivery means of said variable delivery fluid power supply;
a load adaptive energy regenerati g system including first load adaptive energy converting means, said energy accumulating means, and second load adaptive energy converting means; said first load adaptive energy converting means including said energy recuperating pressure drop feedback control system a d operable to convert said load related energy of said fluid motor and load means to a recuperated energy of said energy accumulating means for storing and subsequent use of said recuperated energy; said second load adaptive energy converting means including said supply line pressure drop feedback control system and operable to convert said recuperated energy of said energy accumulating means to a regenerated energy of said pressurized fluid stream being implemented for powering said supply power line of said spool valve means.
38. Λ regenerative adaptive fluid power transmissi n in ~~aT egenerative adaptive fluid motor control system containing a fluid motor control system including fluid motor • means, spool valve means, and fluid power means; said spool valve means having at least three fluid power lines . including a motor line conducting a motor fluid flow to or a motor fluid flow from said fluid motor means a supply power line conducting a supply fluid flow from said fluid power means , and an exhaust power line conducting an exhaust fluid flow to said fluid power meansj said regenerative adaptive fluid power transmission comprising: an exhaust line energy recuperating means for varying a counterpressure rate in said exhaust power line and for recuperating an exhaust fluid energy of said exhaust fluid flow; an energy recuperating pressure drop feedback control system including said.'exhaust line energy recuperating means and operable to regulate an exhaust fluid pressure drpp across said spool valve means by varying said counterpressure rate in said exhaust power line by said exhaust line energy recuperating means;
a variable delivery fluid power supply being powered by energy accumulating means and generating a pressurized fluid stream being implemented for powering said supply power line of said spool valve means;
a supply line pressurfe driop feedback control system including said variable delivery fluid power supply . and operable to regulate a supply fluid pressure drop across said spool valve means by varying a pressure rate of said pressurized fluid stream by a variable delivery means of said variable delivery fltøid power supply;
a load adaptive energy regenerating system including first load adaptive energy converting means, said energy, accumulating means, and second load adaptive energy converting means; said first load adaptive energy converting means including said energy recuperating pressure drop feedback control system and operable to convert said exhaust fluid energy of said exhaust fluid flow to a recuperated energy of said energy accumulating means for storing and subsequent use of said recuperated energy; said second load adaptive energy converting means including said supply line pressure drop feedback control system and operable, to convert said recuperated energy of said energy accumulating means to a regenerated energy of said pressurized fluid stream being implemented for powering said supply power line of said spool valve means,.
39. A regenerative adaptive vehicle drive system comprising: a fluid motor control system including fluid motor and load means, spool valve means, and fluid power means j. said fluid motor and load means including fluid motor means and a mass of a wheeled vehicle and accumulating a mechanicn. energy of said mass of said wheeled vehicle; said spool valve means having at least four fluid poWer lines first ihcludihg~avπϊδTor line conducting a motor fluid flow to said fluid motor means, a second motor line conducting a motor fluid flow from Said fluid motor means, a supply power line, and an exhaust power line» said fluid power means including energy accumula ing means being implemented for powering a variable delivery fluid power supply generating a pressurised fluid stream being implemented for powering said supply power line of said spool valve means; an exhaust line energy recuperating means for varying a counterpressure rate in said exhaust power line and for recuperating said mechanical energy of said mass of said wheeled vehicle; an energy recuperating pressure drop feedback control system including said exhaust line energy recuperating means and operable to regulate an exhaust fluid pressure drop across said spool valve means by varying said counterpressure rate in said exhaust power line by said exhaust line energy recuperating means; a supply line pressure drop feedback control system including said variable delivery fluid power supply and operable to regulate a supply fluid pressure drop across said spool valve means by varying a pressure rate of said pressurized fluid stream by a variable delivery means of said variable delivery fluid power supply;
a load adaptive energy regenerating sy(atem including first load adaptive energy converting means, said energy accumulating means, and second load adaptive energy converting means;
Bai first ioad adaptive energy converting means Including said ehergy recuperating pressure drop feedback control system and operable to convert said mechanical energy of said mass of said wheeled vehicle to a recuperated energy of said energy accumulating means for storing and subsequent use of said recuperated energy; said second load adaptive energy converting means including said supply line pressure drop feedback control system and operable, to convert said recuperated energy of said energy accumulating means to a regenerated energy of said pressurized fluid stream being implemented for powering said supply power line of said spool valve means. , . ,
126
40. The drive system according, to claim 39 wherein said variable delivery fluid power supp y includes a variable displacement pump generating said pressurized fluid stream, wherein said fluid power means include a primary engine being implemented for driving said variable displacement pump, wherein said energy accumulating means are implemented for powering said variable displacement pump, and wherein varying said pressure rate of said pressurized fluid stream is accomplished by a variable displacement means of said variable displacement pump.
The drive system according to claim 39
wherein said variable delivery fluid power supply includes a variable displacement pump generating said pressurized fluid stream, wherein said variable displacement pump is driven by hybrid motor means including an engine, an energy regenerating fluid motor, and an electrical motor, wherein said energy accumulating means include fluid energy accumulating means being implemented for powering electrical energy accumulating means, wherein said fluid energy accumulating means are implemented for powering said energy regenerating fluid motor, wherein said electrical energy accumulating means are implemented for powering said electrical motor, and wherein varying said pressure rate of said pressurized fluid stream is accomplished by a variable displacement means of said variable displacement pump.
42. The drive system according to claim 39 wherein said fluid motor means include a variable displacement motor, and wherein said drive system further comprising: a displacement feedback control system including a variable displacement mechanism of said variable displacement motor and operable to regulate a mechanism displacement of said variable displacement mechanism at least approximately in accordance with a mechanism displacement command signal being correlated with a spool displacement signal of said spool valve means.
43 The drive system according to claim 39 wherein said energy accumulating means include a hydraulic accumulator, pnd wherein said drive system further comprising! a recuperative braking variable displacement pump recuperating said mechanical energy of said mass of said wheeled vehicle and being implemented for powering said hydraulic accumulatort a displacement . control system including a variable displacement mechanism of said recuperative braking variable displacement pump and operable to regulate a mechanism displacement of said variable displacement mechanism at least approximately in accordance with a pressure rate of said motor fluid flow from said fluid motor means ; supplementary energy converting means including said recuperat ve braking variable displacement pump operable to and said displacement control system and Vconvert said mechanical energy of said mass of said wheeled vehicle to said recuperated energy of said energy accumulating means including said hydraulic accumulator, load adaptive in order to assist said firs 'energy converting means-
44. A regenerative adaptive fluid motor output feedback control syste.m comprising: a luid .motor output feedback control system including fluid motor and load means, spool valve means, fluid power means, and output feedback control means; said fluid motor. and load means including fluid motor means and motor load, means, and accumulating a load related energy •• said spool al e means having at least three fluid power liΛβa . Including a motor iine conducting a motor fluid flow to or a motor fluid flow from said fluid motor means of said fluid motor and load means, a supply power line, and an exhaust power line; said fluid power means including energy accumulating means being implemented for powering a variable delivery fluid power supply generating a pressurized fluid stream being implemented for powering said supply power line of said spool valve means;
said output eebac control means measuring a motor outpu of βaid, fluid motor tnβαna and- producing an output feedback control error βignal being implemented for modulating "βaid spool valve meanu;
an exhaust line energy recuperating means for varying a counterpressure rate in said exhaust power line and for recuperating said load..related energy of said fluid motor and load means; rate in said exhaust power line, y
a load adaptive energy regenerating system including first load adaptive energy converting means, said energy, accumulating means, and second load adaptive energy converting meanβ; said first load adaptive energy converting means including said energy recuperating pressure drop feedback control system and operable to convert said load related energy of said fJ.ui< motor and load means to a recuperated energy of said energy accumulating means for storing and subsequent use of said recuperated energyj said second load adaptive energy converting means including said supply line pressure drop feedback control system and operable, to- convert said recuperated energy of said energy accumulating means to a regenerated energy of said pressurized fluid stream being implemented for powering said supply power line of said spool valve means,.
45. The- system according to claim ' 44 , ' herei said fluid motor output feedback control syste is represented by a fluid mot'or position feedback control system, wherein said output .feedback control means are represented by positάon 'feedbook control weans , wherein said motor output is represented by a motor osition, ".■' •' and wherein said output feedback control error signal is represented by a position- feedback control error signal •
46. A regenerative adaptive press drive system comprisin :
- a luid motor position feedback control system including fluid motor and load means, spool valve means, fluid power means, and position feedback control means; ^ said fluid motor and load means including luid motor means of a hydraulic press and accumulating a compressed fluid energy of said fluid motor means of said hydraulic press: said spool valve means having at least three fluid power lines including a motor line conducting a motor fluid flow to or a motor fluid flow from said fluid motor πieahs of said hydraulic press, a supply power line, and an exhaust power line; said fluid power meaniβ including energy accumulating means being implemented for powering a variable delivery' [fluid power Bupply generating a pressurized fluid stream being implemented for powering said Bupply power line of said spool valve means; said position feedback: control means measuring a motor posi'tion of said fluid mototf means-.:and-producing a position feedback control erro signSi^being implemented for modulating s'&'icl s pool lve es s•
an exhaust line energy recuperating means for varying a counterpressure rate in said exhaust power line and for recuperating said compressed fluid energy of said flu±d^mύfό?ή(B§fis <fvsaid hydraulic press; an energy recuperating pressure drop- feedback control system including said exhaust line energy recuperating means and operable to regulate an exhaust fluid pressure drop across said spool valve means .by varying said counterpressure rate in said exhaust power line by said exhaust line energy .recuperating means; said energy recuperating pressure drop feedback control system having a significant dynamic performance superiority against said fluid motor position feedback control system by having either a significant frequency-response superiority or a significant final-transient-tiπie superiority against said fluid motor position feedback control system, a, supply line pressure dπj>p feedback control system including said; ariable delivery luid power supply and operable to regulate a supply fluid presfcjijire drop across said spool valve meanp by varying a pressure rate of said pressurized fluid stream by a variable delivery means of said variable delivery fluid power supply; said supply line pressure drop feedback oontrol system having a significant dynamio./p rformanoe superiority against said fluid motor position feedback control system by having either a significant frequency-response superiority or a significant final-transien n'time superiority against said fluid motor position feedback- control system; a load adaptive energy regenerating system .including first load adaptive energy converting means, said energy, accumulating means, and second load adaptive energy converting means; said first load adaptive energy converting means including said energy recuperating pressure drop feedback control system and operable to convert said compressed fluid energy of said fluid motor means of said hydraulic press to a recuperated energy of said energy accumulating means for storing and subsequent use of said recuperated energy; said second load adaptive energy converting means including said supply line pressure drop feedback control system and operable to convert said recuperated energy of said energy accumulating means to a regenerated energy of said pressurized fluid stream being implemented for powering said supply power line of said spool valve means.
47. The drive system according to claim 46 , wherein said fluid motor means include at least~~bήe hydraulic cylinder having a loadable chamber being loaded against a frame of said hydraulic press, and wherein said compressed fluid energy of said fluid motor means of said hydraulic press includes a compressed fluid energy of said loadable chamber of sai -hydraulic cylinder of said hydraulic press-
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