Patent Publication Number: US-2009220352-A1

Title: Method and Device for Monitoring and Controlling a Hydraulic Actuated Process

Description:
INTRODUCTION 
     1. Field of the Invention 
     The present invention relates to a method for monitoring and controlling a hydraulic pump. More specifically, the invention relates to a method and device for monitoring and controlling a hydraulic actuated process indirectly by monitoring and controlling an electric motor driving a positive displacement hydraulic pump. 
     This invention also relates to a precision hydraulic energy delivery system. Direct coupling of the pump to a primary mover (motor) and related motor control allows for complete motion control of a hydraulically driven machine without the use of any downstream devices. By employing motion control algorithms in the motor control, the hydraulic output at the pump head is controlled in a feed forward method. 
     2. Background of the Invention 
     In the prior art, it is well known that in situations where higher pressures of fluid movement are desired, a positive displacement pump is commonly used. A positive displacement pump is usually a variation of a reciprocating piston and a cylinder, of which the flow is controlled by some sort of valving. Reciprocal machinery, however can be less attractive to use than rotary machinery because the output of a reciprocal machine is cyclic, where the cylinder alternatively pumps or fills, therefore there are breaks in the output. This disadvantage can be overcome to a certain extent by: using multiple cylinders; bypassing the pump output through flow accumulators, attenuators, dampers; or waste gating the excess pressure thereby removing the high pressure output of the flow. 
     In addition to uneven pressure and flow output, reciprocating pumps have the disadvantage of uneven power input proportional to their output. This causes excessive wear and tear on the apparatus, and is inefficient because the pump drive must be sized for the high torque required when the position of the pump connecting rod or cam, in the case of an axial (wobble plate) pump, is at an angular displacement versus the crankarm dimension during the compression stroke that would result in the highest required input shaft torque. 
     Moreover, if the demand of the application varies, complicated bypass, recirculation, or waste gate systems must be used to keep the system from “dead-heading.” That is, if flow output is blocked when the pump is in operation, the pump will either breakdown by the increased pressure or stall. If stalling occurs, a conventional induction electric motor will burn out as it assimilates a locked rotor condition with full rated voltage and amperage applied. Typically systems with fixed displacement pumps use a relief valve to control the maximum system pressure when under load. Therefore, the pump delivers full flow at full pressure regardless of the application thus wasting a large amount of power. 
     In this regard, certain prior art that attempts to correct the problems associated with torque output of a pump motor should be noted. 
     In U.S. Pat. No. 5,971,721, an eccentric transmission transmits a torque demand from a reciprocating pump, which varies with time, to the drive motor such that the torque demand on the drive motor is substantially constant. The result is the leveling of torque variation required to drive a positive displacement pump at the transmission input shaft with the effect of constant pump output pressure. This is accomplished by means of eccentric pitch circle sprocket sets with gear belts or eccentric pitch circle matched gear sets. 
     The use of an eccentric gear or sprocket set has a significant effect on the overall torque requirement and the magnitude of the discharge pulse of the pump. But, because most pumps are of a multi-cylinder or are vane or gear types, the pump input shaft torque requirement would not be perfectly counter-acted (leveled) by using the reduction pattern developed by eccentrically matched transmission components. 
     In U.S. Pat. No. 5,947,693, a position sensor outputs a signal by sensing the position of a piston in a linear compressor. A controller receives the position signal and sends a control signal to control directional motion output from a linear motor. 
     In U.S. Pat. No. 4,726,738, eighteen or nineteen torque leads are measured along the main shaft in order to maintain constant shaft velocity revolution and are translated to a required motor torque for particular angles of the main shaft. 
     U.S. Pat. No. 4,971,522 uses a cyclic lead transducer input and tachometer signal input to a controller to signal varied cyclic motor input controls to provide the required motor torque output. A flywheel is coupled to the motor in order to maintain shaft velocity. However, the speed of the motor is widely varied and the torque is varied to a smaller extent. 
     U.S. Pat. No. 5,141,402 discloses an electrical current and frequency applied to the motor which are varied according to fluid pressure and flow signals from the pump. 
     U.S. Pat. No. 5,295,737 discloses a motor output which is varied by a current regulator according to a predetermined cyclic pressure output requirement. The motor speed is set to be proportional to the volume consumed and inversely proportional to the pressure. 
     It is seen from the foregoing that there is a need for electronic attenuation of the torque profile in a pump. When the torque profile is compared with the input shaft displacement and other known factors such as system inertia and response time of the pump drive etc., a pump can produce constant pressure and therefore constant flow without the typically associated ripple common to power pumps for the full range of the designed volumetric delivery, by driving them in a feed forward method. 
     It should be noted that the foregoing hydraulic pumping systems control output pressure and flow in the micro sense. These concepts examine modulating the input shaft torque and speed to provide a constant hydraulic output, whether it is pressure or flow limited. See U.S. Pat. No. 5,971,721, U.S. Pat. No. 6,494,685, and U.S. Pat. No. 6,652,239, the entire contents of which are hereby incorporated by reference. 
     It should be further noted that attempts to provide a high dynamic range of hydraulic flow and pressure during operation of prior pumping systems required placement of downstream devices in the liquid path to modulate the hydraulic output. With such systems, the pump provides the maximum hydraulic flow (as the prime mover) and the downstream devices adjust the output to match the application requirements. 
     The prime mover in such systems is typically a constant speed induction motor. In to order to control the hydraulic output, feedback devices, a processor (be it mechanically balanced or electronic) and hydraulic servo valves must be placed into the hydraulic stream for flow and pressure regulation. This treatment of hydraulic delivery places the “smarts” of the system in the hydraulic output portion of the system. Disadvantageously, these systems require many hydraulically driven devices, are mechanically (geometry) limited, are energy inefficient when total system performance is scrutinized and have a small range of dynamic response (typically 10-1). Typical examples of a command-response curve of a small servo valve and a large servo valve are shown in  FIGS. 9 and 10 , respectively. 
     Moving the “smarts” directly into the prime mover—by incorporating variable speed (VFC) controlled motors—has been attempted. However, this provides limited torque delivery potential at low speeds, and many feedback devices are required for its operation. Further, the response of such a system is only generally higher than the 150 ms range and the energy savings potential is only in the 50% range. 
     These approaches address—in the macro sense—the need for a prime mover coupled to a power pump that controls the energy, and therefore the flow (velocity) and pressure (torque) at the input shaft of the pump. Moreover, the desired system must replicate the motion control capabilities of existing systems without requiring the use of downstream flow control devices and feedback circuits. 
     An example of another hydraulic process, also to which one embodiment of the present invention is directed to, would be presses (such a metal forging presses and laminating presses) driven by hydraulic cylinders. In critical applications, valves and feedback controls are used to control the critical process variables of applied force and the rate of actuation. A variety of transducers are employed both to facilitate control and to monitor the process. 
     Additionally, level switches, pressure transducers and flow switches are often employed to monitor system conditions such as low oil levels, plugged pump inlet filters and burst hoses. Use of such transducers adds to the system cost and complexity and is prone to failure particularly in hostile environments. 
     Yet another example of a prior art implementation are computer controlled servo valves relying on feedback from pressure transducers, linear position transducers, and flow meters. The limitations of traditional systems include reliability, slow response time and control precision. A recent research paper in the Journal of Materials titled:  Modeling and Simulating Metal - Forming Equipment,  by Frazier, Medina, Mullins and Irwin describes the current art as it is applied to forging presses. The paper described a metal forging process, which included a cylindrical upsetting of plain carbon steel. The press was programmed to forge at a constant velocity of 1.27 cm/sec. A plot of an experimental and simulated ram velocity profile is shown in  FIG. 12 . The data clearly reveals the ramping up and overshoot of the desired velocity. A brief change in velocity due to impact with the workpiece is observed near 3.5 seconds. 
       FIG. 13  shows an experimental and simulated result for the ram load as derived from the ram pressure measurement. The plot reveals that approximately 88.96 kN are needed to overcome the counter-balance and frictional forces. The load increases rapidly beginning at approximately 3.5 seconds. This corresponds to impact with the workpiece. Beginning at approximately 4 seconds, elastic deformation of the workpiece and tooling ceases and plastic deformation of the workpiece begins. It is clear from the data that the press was able to maintain the desired velocity under load, as long as the load did not increase too rapidly, but as soon there was a sudden increase in load, the desired velocity is lost. 
     It is to be noted that a particular valve command does not always result in the same flow rate. This is due to the changing pressure across the servo manifold, clearly revealing that the flow rate can decrease even as the command for higher velocity increases. 
     As the need for precise control of ram velocity, as well as position, increases, the need for direct measurement of velocity as a feedback control signal will become more acute. Therefore, current techniques for controlling velocity (inversion of servovalve flow models combined with pressure measurements and numerical differentiation of displacement measurements) are not adequate for high performance over a broad range of forging conditions. 
     Experience shows the need for press operators to customize the press control law in order to achieve the desired velocity profile for different forming operations. Having a custom control law for equipment that repeatedly makes the same part for several weeks or more is satisfactory, but as the need to use the same equipment for several different parts in a day or custom small lots increases, having to repeatedly tune the control law can waste significant time. Control laws, therefore, need to be designed to be robust so that different loading conditions and velocity profiles can be handled successfully without the need for customized tuning. 
     The paper&#39;s conclusion includes the following statement:
         Current techniques for controlling velocity (inversion of servovalve flow models combined with pressure measurements and numerical differentiation of displacement measurements) are not adequate for high performance over a broad range of forging conditions.       

     Accordingly, there is a need for a system and method of eliminating the need for servovalves and transducers while providing more precise control of both hydraulic press ram force and displacement. There is also a need for a pump and motor assembly with constant pressure output and a motor controller with electrical regeneration. With the possibility of directly sensing workpiece conditions during forming operations, it is demonstrated that these measurements could be fed back to the metal-forming equipment for computer control of the forming equipment, thereby enabling real-time compensation for variations in initial workpiece and equipment conditions. From a process control perspective, this approach enables the highest level of robustness and repeatability in production. Common to each of these factors is the need for improved control and predictability of the equipment&#39;s behavior, which itself is based on the desire to achieve near-net shapes, higher-quality end products, higher yields, and better control of microstructure, especially for hard-to-form materials in metal forging process. 
     More specifically, the current synchronized motion control methods used in hydraulic applications are not designed to possess the dynamic hydraulic properties which affect each axis independently that are responsible for coordinating force inherent in classical multi-axis, hydraulic machinery, such as for example shown in  FIG. 15 . Consequently, these controllers cannot easily maintain coordination for all operating conditions without expensive and complicated feedback devices and control loops. 
     SUMMARY OF THE INVENTION 
     It is therefore an object of the invention to device a control method for use in hydraulic applications that possesses the dynamic hydraulic properties affecting each axis independently that are responsible for coordinating force inherent in classical multi-axis and multi-prime mover, hydraulic machinery. 
     A further objective of the invention is to provide controllers that maintain coordination for all operating conditions without expensive and complicated feedback devices and control loops. 
     Another object of the present invention is to provide a method for electronic attenuation of pump torque variation requirements in order to produce a matched motor torque output that will result in constant output pressure from a pump. 
     Yet another object of the present invention is to provide monitoring and control factors which vary the power and torque output of a pump motor based on calculated torque variation requirements. 
     Yet another object of the present invention is to increase the energy efficiency of a pump system, by providing a force balanced relationship between the motor output and the application&#39;s hydraulic requirement, thus allowing the use of energy saving torque drives without incurring the pressure variations associated with their use. 
     Yet another object of the present invention is to decrease the wear and tear on the pump by providing a substantially constant force output from the motor of the pump and reduce the amount of cycles of the pump to the application&#39;s requirement. 
     Yet another object of the present invention is to provide a method for electronic attenuation of pump torque variation by supplying information for design of an electronic transmission system that can achieve a modulated torque output from the motor to the pump. 
     Yet another object of the present invention is to achieve precise control of flows and pressures, thus precise control of press ram velocities, force and position. 
     Yet another object of the invention is to create precise hydraulic output from the pump utilizing an algorithm programmed into the drive control that systematically measures and corrects for three key physical parameters of the motor, pump and hydraulic fluid combination. The three parameters being windage torque, viscous torque and coulomb torque. One advantage provided by this aspect of the invention is that the resultant precision allows multiple (two or more) drive, motor and pump combinations to be electronically “line-shafted” together while providing precise and stable control of pressure and flow being fed to a common output header regardless of individual pump characteristics or fluid condition variations. 
     The invention also provides for development of an empirical understanding of the positive displacement pump in regards to slippage (variation from theoretical displacement) throughout the full pressure/flow delivery range of the pump. This understanding combined with the drive controller&#39;s precise measurement of motor rpm gives an extremely precise measurement and control of hydraulic fluid flow rate. 
     In turn, knowing the displacement volumes of the cylinder being actuated and mathematically correcting for elasticity of the piping system allows for very precise control and or measurement of press-ram position, and/or ram velocity than may be achieved with prior art. 
     The invention employs the feed forward torque control aspects of the drive system to precisely monitor and/or control hydraulic pressures created by the pump which after mathematical correction for line losses at various flow rates, mimics load at the press ram. The invention also provides both a more precise and a more robust method than does the prior art. 
     To attain the objects described, there is provided a method for obtaining a polar map for process control within the electronic drive of a targeted pump. This polar map is calculated by a processor or is externally calculated then input into a processor. Once the torque profile of the pump is obtained and translated into a polar map, the processor can compare the shaft displacement angle of the pump input shaft to the reference polar map. The processor can also take into account selected factors such as the response time of the pump drive, the motor inductive reactance, system inertia, application characteristics of the pump, and regenerative energy during deceleration of the pump. 
     Using selected factors and the comparison results, the processor signals the motor controller to vary the amperage, voltage, and frequency applied to the motor in order to regulate the torque output of the pump motor. With an accurately modulated motor torque output in concert with the established polar map (for the targeted pump), the pump output pressure will remain constant regardless of the pump&#39;s crank arm location or the velocity of fluid flow. 
     It is also an object of the present invention to provide a hydraulic energy delivery system that allows for complete motion control of a hydraulically driven machine with the use of minimal or no downstream feedback devices. 
     It is therefore a further object of the present invention to provide control factors which vary the power and torque output of a pump motor by employing motion control algorithms. 
     To attain the objects described, there is provided direct coupling of a positive displacement pump to a pump drive motor and related controls. By employing motion control algorithms into the motor control, the hydraulic output at the pump head will simultaneously follow. Control features listed herein may be integrated into the system by developing algorithms and subroutines for the control system coupled to the pump. 
     These and other objects and advantages are provided by the present invention, preferred embodiments of which are described in the paragraphs below. 
     All features listed herein may be integrated into the system by developing algorithms and subroutines for monitoring and controlling the system coupled to the hydraulic pump. 
     The present invention will now be described in more complete detail with reference being made to the figures identified below. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       The following detailed description, given by way of example and not intended to limit the present invention solely thereto, will best be appreciated in conjunction with the accompanying drawings, wherein like reference numerals denote like elements and parts, in which: 
         FIG. 1  is a block diagram of the steps required for a method of electronic attenuation of torque profile and the resulting control of the pump; 
         FIG. 2  is a graph depicting individual input torque variation for each node of a triplex pump based upon pump input shaft rotational degrees; 
         FIG. 3  is a graph depicting a percentile summation of input torque variation compared to angular displacement of the input shaft of a triplex pump; 
         FIG. 4  is a table depicting variations of input torque above and below the mean for triplex pumps in relation to the linear distance between the plunger/piston pivot point and the throw pivot point multiplied by the throw radius; 
         FIG. 5  is a graph depicting a plotting of geometric distance variation points based upon the summation of total torque variation for a triplex pump; 
         FIG. 6  is a polar map depicting the torque profile versus angular displacement of a pump input shaft; 
         FIG. 7  is a diagram illustrating a precision hydraulic delivery system according to the present invention; 
         FIG. 8  is a graph depicting a profile of torque vs. velocity for an exemplary hydraulic system in accordance with the present invention; 
         FIG. 9  is a graph depicting command, response curves of a small servo valve; 
         FIG. 10  is a graph depicting command, response curves of a large servo valve; 
         FIG. 11  is a schematic of a hydraulic press system according to one embodiment of the invention; 
         FIG. 12  is a graph depicting commanded, simulated and experimental ram velocity profiles not within normal boundaries; 
         FIG. 13  is a graph depicting simulated and experimental ram load profiles not within normal boundaries; 
         FIG. 14  is a diagram of “electronic line-shafting” with a master/slave configuration of multi-drop positive displacement hydraulic pumps, according to one embodiment of the invention; 
         FIG. 15  is a schematic of a typical non-compensated electronic line-shafting in a multi-axis hydraulic machinery; and 
         FIG. 16  is a diagram of a system incorporating the Learn TQ items: Coulomb torque, Windage, viscous coefficient and application characteristics: Redux P, Redux V feed-forward compensation algorithm, according to one embodiment of the invention. 
     
    
    
     The description of the various elements of the invention will be discussed in detail in the following sections. 
     DETAILED DESCRIPTION OF THE INVENTION 
     The instant invention will now be described more fully hereinafter with reference to the accompanying drawings, in which preferred embodiments of the invention are shown. This invention may, however, be embodied in many different forms and should not be construed as limited to the illustrated embodiments set forth herein. Rather, these illustrated embodiments are provided so that this disclosure will be thorough and complete, and will fully convey the scope of the invention to those skilled in the art. 
     Referring now to the drawings in detail wherein like numerals refer to like elements throughout the several views where Blocks  1 - 5  of  FIG. 1  depict the development of a baseline polar guide of the torque profile for the targeted pump. 
     In Block  1  of  FIG. 1  and graphically depicted in  FIG. 2 , the output characteristic of volumetric displacement would directly relate to the input torque variations above  10  and below  12  the comparative mean  14 . The processor identifies the output discharge characteristics such as the number of plungers, pistons in a piston pump, or vane/gear in a rotary pump. The processor also utilizes a comparative mean where, the comparative mean is representative of the basic torque requirement of the pump input shaft rated at a specific output pressure of the pump. A pulsation pattern  16  would be repeated at the same rate per revolution as the number of the pump&#39;s volumetric displacement cavities. As illustrated in  FIG. 2 , a triplex positive displacement pump would repeat a pulsation pattern  16  every 120 degree rotation of the pump input shaft. These torque variations above  10  and below  12  the mean  14  are calculated and recorded for Block  1  of  FIG. 1 . 
     For other pumps such as a quintaplex plunger pump, which incorporates five plungers, a pulsation pattern would be produced five times per revolution of the pump input shaft, repeating every 72 degrees if the output pressure is to remain constant; and for a rotary vane pump with nine vanes selected, the pulsation pattern would repeat every 40 degree rotation of the pump input shaft if the output pressure is to remain constant. 
     In Block  2  of  FIG. 1  and depicted graphically in  FIG. 3 , the torque profile versus displacement angle of the targeted pumping system is the summation of the torque requirement for each volumetric displacement component, depicting a percentage above mean  18  and the percentage below mean  20 . 
     In Block  3  of  FIG. 1 , the magnitude of the input torque variation for the power pump is determined by the processor, where the magnitude of the torque variation is the number of volumetric displacement cavities activated in one revolution and the relationship “Q”. The calculation “Q” is the linear distance “L” between the plunger/piston pivot point and the throw pivot point multiplied by the throw radius “R”; “Q=LR”.  FIG. 4  in table form, depicts the percentile variations of input torque above and below the mean for triplex pumps with various “Q”. 
       FIG. 5  graphically depicts the total torque variation to show a torque profile for a triplex pump (three volumetric displacements per revolution) with a “Q” at 4:1 with variations shown above and below the mean. The mean is representative of the basic rms (root mean squared) torque requirement of the pump input shaft rated at a specific output pressure of the pump versus the angular displacement of the pump crank shaft. The relationship of “Q” and the effect it has on torque variation would also apply to rotary pumps. A plotted geometric distance variation using t 1 -t 15  (as plotting points) is then imposed on the torque profile. 
     In Block  4  of  FIG. 1  and graphically depicted in  FIG. 6 , a pump polar map is determined based on the torque profile and the input shaft angular displacement of the pump. The center  34  of the polar map is to represent zero torque. The incremental lines  36  depicted orbitally are the angular displacement of the targeted pump&#39;s input shaft. The plotted pump torque variation curve  38  that occurs above and below the mean  40  is to be considered a geometric percentage of the summation of the torque requirement of each of the volumetric displacement components of the targeted pump. 
     The distance of each point plotted on the polar map&#39;s center from the base diameter&#39;s center is the geometric distance variation (over or under) of the base radii percentile established from torque versus the pump input shaft displacement angle (t 1  thru t 15 ). The geometric distance variations are the plotting points determined in  FIG. 5 . The torque versus angular displacement profile of the pump system selected is to become the reference polar guide for the comparitor algorithm in the processor in Block  5  of  FIG. 1 . The reference polar guide determined by the processor in Blocks  1 - 5  can also be determined externally from the processor and then input into the processor. 
     Blocks  6 - 10  of  FIG. 1  are the operating steps from electronic attenuation of the torque profile to provide a constant output pressure at the pump, wherein Block  6  indicates the transmission of the angular displacement of the input shaft of a pump in operation. A pulse transmitter mounted on the input shaft relays to a counter—which is part of the processor—the angular position of the pump drive. 
     In Block  7  of  FIG. 1 , an electronic processor gathers this output shaft orientation feedback information, and processes the angular displacement data. The processor then attenuates from the peak requirement of the pump, the output torque of the drive compared to the predetermined reference polar map of Block  5 . A corresponding torque command value is then selected. 
     In Block  8  of  FIG. 1 , other inputs of system readings such as system inertia, parasitic leads, off throttle friction, response time of the pump, motor inductive reactance, application characteristics of the pump, regenerative energy during deceleration of the pump, and translation speed can be selectively factored into the processor algorithm for changes in process control. 
     In Block  9  of  FIG. 1 , based upon the inputs of Blocks  7  and  8 , the processor of the electronic drive signals the motor controller to apply the correct amperage, voltage, and frequency to the motor which then provides the correct torque according to the angular displacement of the pump input shaft. 
     In Block  10  of  FIG. 1 , the resultant signal to the motor controller and motor will drive the pumping system to produce constant pressure at the full range of the designed system flow volume regardless of pump radial crankshaft location and the velocity of the fluid pumped. 
     Block  11  of  FIG. 1 , depicts the use of this method in future systems where information gathered from pump operation by this method can be used to design more responsive components such as transmissions and electronic drives. More responsive components would decrease the time increments between Blocks  6 - 10 . As response times are decreased, the torque output produced for indicated angular displacements will increase in efficiency. 
       FIG. 7  depicts a precision hydraulic delivery system  71  according to the present invention. Advantageously, this system provides direct coupling of a positive displacement pump  72  to a prime mover  73  and related motor drive control  74 . The prime mover  73  in the pump system shown is, for example, a constant speed induction motor. The motor has, for example, a 1000-1 (torque) turn down ratio. The motor control  74  may be, for example, an electronic servo or hydraulic type motor control. Direct coupling of the pump  72  to the motor  73  and motor control  74  allows for complete motion control of the pump  72  without requiring any of the downstream flow control devices, feedback devices, hydraulic energy storage devices (accumulators) or energy dissipation devices normally used in conventional pump systems. 
     The system in  FIG. 7  employs motion control algorithms in the electronic motor control so that the hydraulic output at the pump head will simultaneously follow the control signals generated by the algorithms and sent to the motor. This ability allows a large dynamic range of hydraulic energy to be delivered by placing the “smarts” of the system directly into the electrical handling capabilities of the prime mover circuit. The modulation of torque (resulting in hydraulic pressure) and velocity (resulting in hydraulic flow) are most efficiently handled within the electronic servo or hydraulic type control of the primary mover. 
     The teachings of U.S. patent application Ser. No. 09/821,603 and U.S. Pat. No. 5,971,721, which are hereby incorporated by reference, may be incorporated into the macro motion control capabilities described herein to provide improved system response, “keypad” tuning of a hydraulic application, very high systemic efficiency characteristics and simplified hydraulic circuitry. 
     Several exemplary control features of the present invention are described in greater detail below. These features represent only a fraction of the possible features that may be electronically integrated into a hydraulic delivery system by control algorithms and subroutines for a prime mover servo control system coupled to a pump. 
     “SLAM Absorption” Feature 
     The “SLAM” subroutine is an energy absorbing function that provides hydraulic component protection by eliminating pressure spikes. In some applications, a “spike” in pressure occurs when flow volume is rapidly reduced. This normally occurs when, for example, a directional control valve is shut, and is typically followed by the pressure relief valve waste-gating the excess flow to a tank until the system flow returns to normal. 
     This condition is undesirable, and to eliminate it the present invention has a discrete input that activates the “SLAM” function when such an event occurs. A determination as to the likelihood of such an event is made during commissioning. Use of the “Position Sensing” feature (described below) allows the “SLAM” subroutine to be invoked when necessary. The “SLAM” feature causes the electronic drive to capture the inertial energy of the system via the regenerating capabilities of the prime mover (turning the motor into a generator), and to store this captured electrical energy in the electronic drive (see “energy storage system” below). The normally waste-gated energy is thus captured by the drive during this function, thereby saving energy and reducing wear on the hoses and hydraulic system. 
     “JAB Applied” Feature 
     The “JAB” feature eliminates pressure “droop” by invoking a rapid pump acceleration feature of user defined time and amplitude, that is applied over and above the normal flow or pressure input commands. In some instances, a rapid increase in flow volume required by the application will cause the pressure to droop until high inertia components in the pumping system are accelerated to the required delivery velocity. If this droop is undesirable in a specific application, a discrete input can be used to activate this “JAB” rapid acceleration feature that is applied over and above the normal flow or pressure input commands that are controlling the pump. 
     Dual Function Pump/Motor Feature 
     This feature provides for single unit hydraulic motor/pump functions from the same hydraulic device for energy delivery and reclamation (regeneration and storage). 
     “Pressure Loop” Feature 
     This feature provides a pump shaft torque output measurement method which is translated into a pressure delivered signal. 
     “Constant HP System” Feature 
     This feature provides a constant horse power electrical drive system for maintaining an energy ceiling regardless of the delivered flow volume. 
     “Energy Storage System” Feature 
     This feature provides an electrical energy storage device in the drive system for reclamation of energy from regeneration (see “Dual function pump/motor” and “SLAM” function), or for high output energy spikes typically provided by a hydraulic accumulator. “Position Sensing” feature 
     According to this feature, a volumetric pulse correlates to a pump output volume that will cause an incremental pulse to occur. This volumetric pulse (output by the electronic drive module inclusive of compensation factors Pump TQ Learn and application items REDUX V and REDUX P) is used for the positioning of known hydraulic cylinders and their corresponding volumetric displacements. 
     “Leakage Detection” Feature 
     This subroutine is used to detect user defined excessive hydraulic leakage rates. This feature compares the output of the “Position Sensing” function to a known limit during a move, and if there is a discrepancy beyond a predetermined amount, an alarm output results. 
     “Output Gain Offset” Feature 
     This feature allows the user to assess the output gain levels of the hydraulic delivery (pressure vs. flow) in order to overcome any application flow restrictions or mechanical variation. The assessment results in a profile of torque vs. velocity for the desired hydraulic output. 
       FIG. 8  shows an example 5 point torque profile, including:(1) Gain Zero 801, (2) Gain Lo  802 , (3) Gain Mid  803 , (4) Gain Hi  804 , and (5) Gain Max  805 . The five gain points plotted on the graph are described below. 
     1. Gain Zero: For “pressure delivered” vs. “zero velocity” (the RPM of this point is always anchored at zero RPM), the Gain Zero corrects the pressure reference command as the velocity decreases to “0” to compensate for systemic “sticktion”. 
     2. Gain Low: For “pressure delivered” vs. “velocity,” the Gain Low corrects the pressure reference command as the velocity increases/decreases to compensate for system losses. Gain Low RPM: Applies the “GAIN LOW” value when the pump system is operating within a user defined RPM range (typically, 0 to 50 RPM). The gain is applied as a tapered offset beginning with the “GAIN ZERO” value at 0 RPM, and ending with the “GAIN LOW” value at the “GAIN LOW RPM.” Any operation above this speed is ramped to the “GAIN MID” point. 
     3. Gain Mid: For “pressure delivered” vs. “velocity,” the Gain Mid corrects the pressure reference command as the velocity increases/decreases to compensate for system losses. 
     Gain Mid RPM: Applies the “GAIN MID” value when the pump system is operating within a user defined RPM range (typically, 50 to 700 RPM). The gain is applied as a continued offset beginning with the “GAIN LO” value at the “GAIN LO RPM” and ending with the “GAIN MID” value at the “GAIN MID RPM.” Any operation above this speed is ramped to the “GAIN HI” point. 
     4. Gain High: For “pressure delivered” vs. “velocity,” the Gain High corrects the pressure reference command as the velocity increases/decreases to compensate for system losses. 
     Gain High RPM: Applies the “GAIN HIGH” value when the pump system is operating within a user defined RPM range (typically,  701  to the maximum RPM). The gain is applied as a continued offset beginning with the “GAIN MID” value at the “GAIN MID RPM” and ending with the “GAIN HIGH” value at the “GAIN HIGH RPM.” Any operation above this speed is ramped to the GAIN MAX RPM point. 
     5. Gain Max: For pressure delivered vs. DRIVE SPEED MAX velocity (the RPM of this point is always anchored at the drive speed max RPM), the Gain Max attenuates the pressure reference command as the velocity increases/decreases to compensate for system losses. 
     The invention according to one embodiment of the invention is a method for monitoring and controlling a hydraulic pump driving a metal forging press. A simplified block diagram of a hydraulic press system  110 , for example a metal forging press, is shown in  FIG. 11 . The system  110  is powered by an electric motor that drives a hydraulic pump  118 . Transient demands for high ram speeds are met by an accumulator system  120 . A counter-balance is employed to support and return the main ram  112  to the top of its stroke after the completion of a forging operation. The hydraulic manifold  114  controls the flow of fluid to the main ram cylinder  116  and to the tank  122 . The metal forging can be a cylindrical upsetting of, for example, plain carbon steel. The press can be programmed to forge at any constant velocity, for example at 1.27 cm/sec. 
     One embodiment of the invention is an application of an “electronic line-shafting” control technique which serves to replicate and even improve the historical, hydraulic multi-axis coordinated motion control techniques. This technique incorporates a method of servo-driven hydraulic prime-mover control on multi-axis hydraulic applications with the Learn TQ, Redux V, and Redux P compensation factors incorporated for direct feed forward precise hydraulic output without the need for external feedback devices. Redux P relates to the compensation or reduction or compression of a fluid or system capacitance, while Redux V relates to the restrictive flow of the pump. The result demonstrates that the “electronic line-shafting” technique significantly improves the coordination, robustness, and overall stability of hydraulic power output subjected to realistic physical limitations. 
     Therefore, key aspects of the present invention include precise control of press ram velocities, force and position requires precise control of flows and pressures. To create precise hydraulic output from the pump, the invention utilizes an algorithm programmed into the drive control named Pump Torque Learn (abbreviated Learn TQ) that systematically measures and corrects for three key physical parameters of the motor, pump and hydraulic fluid combination. These three factors are accounted for to understand the relationship between applied torque and resultant pump pressure; a relationship that changes over the range of operating speeds of the pump. These parameters are windage torque, viscous torque and coulomb torque, respectively. A separate advantage provided by this aspect of the invention is that the resultant precision allows multiple (two or more) drive, motor and pump combinations  125  to be electronically “line-shafted” together while providing precise and stable control of pressure and flow being fed to a common output header  130  regardless of individual pump characteristics or fluid condition variations. A multi-drop set-up for positive displacement hydraulic pumps used in “line-shafting” hydraulic supplies, according to one embodiment of the invention is shown in  FIG. 14 . A further example of a system  150  incorporating the Learn TQ, Redux P, Redux V feed forward compensation algorithm according to one embodiment of the invention is shown in  FIG. 16 . 
     The invention also incorporates development of an empirical understanding of the positive displacement pump in regards to slippage (variation from theoretical displacement) throughout the full pressure/flow delivery range of the pump. This understanding combined with the drive controller&#39;s precise measurement of motor rpm gives an extremely precise measurement and control of hydraulic fluid flow rate. In turn, knowing the displacement volumes of the cylinder being actuated and mathematically correcting for elasticity of the piping system allows for very precise control and or measurement of press-ram position, and/or ram velocity than may be achieved with prior art. 
     As explained above, one embodiment of the invention employs the feed forward torque control aspects of the drive system to precisely monitor and/or control hydraulic pressures created by the pump which after mathematical correction for line losses at various flow rates, mimics load at the press ram. The invention therefore provides both a more precise and a more robust method than does prior art. 
     Accordingly, some of the benefits derived from the present invention are: 
     1. More precise control of ram velocity profile than prior art. 
     2. Without relying on external sensors mounted at the process but rather using motor current signature analysis combined with motor shaft encoder data, the drive controller may be taught to recognize, report, and/or take action against process anomalies. Such anomalies would show themselves as velocity (flow rates), current draws and torques (pressures) that are deemed to be outside normal process parameters. Examples include: 
     a) A plugged inlet filter would cause a low current signature combined with a high velocity (motor rpm) than would be expected for the flow and pressure being delivered. 
     b) A burst hydraulic hose would result in an abrupt shift to an abnormally high flow rate. 
     c) Metal being forged that was not within proper metallurgy specifications or a billet not heated to the desired temperature would produce ram speed, displacement and force characteristics that the computer could recognize or simply log as being not within normal boundaries (such as shown in  FIGS. 12 and 13 ). 
     The ability to “electronically line-shaft” two or more pump, motor, and drive systems to feed a common hydraulic output header while maintaining precise and stable control of both total flow and common pressure. 
     Thus, while fundamental novel features of the invention are shown and described and pointed out, it will be understood that various omissions and substitutions and changes in the form and details of the devices illustrated, and in their operation, may be made by those skilled in the art without departing from the spirit of the invention. For example, it is expressly intended that all combinations of those elements and/or method steps which perform substantially the same function in substantially the same way to achieve the same results are within the scope of the invention. Moreover, it should be recognized that structures and/or elements and/or method steps shown and/or described in connection with any disclosed form or embodiment of the invention may be incorporated in another form or embodiment. It is the intention, therefore, to be limited only as indicated by the scope of the claims appended hereto.