HYDRAULIC SYSTEM WITH SERVO DRIVE AND HYDRAULIC LOAD AND CONTROL UNIT FOR THE HYDRAULIC SYSTEM

Abstract

To utilize and protect a mechanical load torque range of a servo drive in combination with a pump, a control unit is given a target system pressure as a reference variable and an actual system pressure as a control variable. An electric motor torque acting on a pump of the servo drive is specified by the control unit to an electric motor of the servo drive, a volume flow at the hydraulic load is generated by the pump, by which a mechanical load torque sets it at the electric motor and the actual system pressure is produced in the hydraulic load via the volume flow. A dynamic system variable of the hydraulic system is transmitted to the limiting unit. The limiting unit limits the motor torque transmitted by the control unit to the electric motor as a function of the value of the system variable

Claims

1. A method for controlling the actual system pressure of a hydraulic load of a hydraulic system, the method comprising: a control unit is given a target system pressure as a reference variable and the actual system pressure as a control variable; the control unit specifies an electric motor torque to an electric motor of a servo drive, which acts on a pump of the servo drive; the pump generates a volume flow at the hydraulic load, by which a mechanical load torque sets in at the electric motor; and the actual system pressure is generated in the hydraulic load via the volume flow, wherein a limiting unit is given a dynamic system variable of the hydraulic system, wherein the limiting unit limits the motor torque transmitted to the electric motor as a function of the value of the system variable, and wherein a calculation unit calculates an estimated load torque using the system variable and transmits it to the limiting unit, which limits the motor torque as a function of the value of the estimated load torque.

2. The method according to claim 1, wherein a minimum load torque threshold, preferably zero, and/or a maximum load torque threshold is specified to the comparison unit and the estimated load torque is transmitted to the comparison unit by the calculation unit, wherein the comparison unit verifies whether the estimated load torque falls below the minimum load torque threshold and/or exceeds the maximum load torque threshold, and in the event of a pending undershoot/overshoot, a signal is sent to the limiting unit, and wherein the limiting unit limits the motor torque upon receiving the signal.

3. The method according to claim 1, wherein the estimated load torque is calculated using a model of the hydraulic system, wherein the motor speed serves as a system variable.

4. The method according to claim 3, wherein the model is described by the formula $M_{\mathrm{last}}=M_{\mathrm{motor}}-\frac{k_0}{2.Math..Math.}.Math.{}_{\mathrm{motor}}-J_{\mathrm{ges}}.Math.{\stackrel{.}{}}_{\mathrm{motor}}$ having the parameters of electric motor torque, moment of inertia of the motor and torque constant.

5. The method according to claim 3, wherein a corrected torque constant (k.sub.v) is determined from the transmission behavior of the drive line and used in the model.

6. The method according to claim 5, wherein the corrected torque constant is calculated at an operating point from the relationship $k_v=\frac{\left(1-{}_{\mathrm{pump}}\right).Math.V_{\mathrm{th}}.Math.p_{\mathrm{ist}}}{20.Math..Math..Math.{}_{\mathrm{pump}}.Math.n},$ using the parameters of pump volume of the pump, actual system pressure, pump efficiency and motor speed.

7. The method according to claim 1, wherein the limiting unit obtains the system variable from the control unit and/or the servo drive and/or the hydraulic load.

8. A hydraulic system comprising: a servo drive composed of an electric motor and a pump; a control unit; and a hydraulic load, wherein the control unit is given a target system pressure as a reference variable and an actual system pressure of the hydraulic load as the control variable, wherein the control unit specifies to the electric motor an electric torque as a variable, wherein the electric motor transmits the motor torque to the pump, whereby the pump generates a volume flow at the hydraulic load, by which the actual system pressure is generated, and wherein a mechanical load torque sets in at the electric motor, wherein a limiting unit is connected to the control unit, wherein the limiting unit limits the motor torque transmitted by the control unit to the electric motor by using a system variable of the hydraulic system, and wherein a calculation unit is on hand, which, by using the system variable, calculates an estimated load torque and transmits it to the limiting unit, which limits the motor torque as a function of the estimated load torque.

9. The hydraulic system according to claim 8, wherein the limiting unit is an integral component of control unit.

10. The hydraulic system according to claim 8, wherein the calculation unit is an integral component of limiting unit.

11. The hydraulic system according to claim 8, wherein a comparison unit, which is connected to the calculation unit and the limiting unit, is present, wherein the comparison unit receives the estimated load torque from the calculation unit and verifies whether the estimated load torque falls below a minimum load torque threshold and/or exceeds a maximum load torque threshold, and in the event an imminent undershoot/overshoot transmits a signal to the limiting unit, which limits the motor torque upon receiving the signal.

12. The hydraulic system according to claim 11, wherein the comparison unit is an integral component of the limiting unit.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

[0015] The present invention is explained in greater detail below with reference to FIGS. 1 to 6, which depict advantageous embodiments of the invention in an illustrative, schematic and non-restricting manner. Thereby depicted are:

[0016] FIG. 1 a hydraulic system having a limiting unit according to the invention,

[0017] FIG. 2 a limiting unit having a comparison unit and a calculation unit,

[0018] FIG. 3 the transmission function of electrical motor torque to the motor speed,

[0019] FIG. 4 the trends of the specified motor constant, calculated motor constant and actual motor constant,

[0020] FIG. 5 the trends of estimated load torques and the actual mechanical load torque,

[0021] FIG. 6 the trends of a load torque, an estimated load torque and a motor torque.

DETAILED DESCRIPTION OF THE EMBODIMENTS

[0022] FIG. 1 depicts a hydraulic system 1. A hydraulic load 5 has an actual system pressure p.sub.ist , which sets in due to a supplied volume flow V of a medium. Volume flow V is generated by a pump 3, which in turn is driven by an electric motor 2, which typically results via a coupling of a motor shaft of electric motor 2 to a pump shaft of pump 3. Electric motor 2 and pump 3 together form a servo drive 9, which thereby pumps the pumped medium, e.g. hydraulic fluid, to the hydraulic load at volume flow V. To this end, an electric torque M.sub.motor is generated at electric motor 2 via a torque-generating current. Mechanical load torque M.sub.last thus represents the torque, which sets in at the inlet of pump 3, e.g., at the pump shaft, when that volume flow V is provided in hydraulic load 5, which produces the desired actual system pressure p.sub.ist.

[0023] A control unit 4, e.g., a programmable logic controller (PLC), is given a target system pressure p.sub.soll as a control variable, wherein this specification may be provided for example by a user or a control program. In addition, control unit 4 also receives current actual system pressure p.sub.ist as a feedback control variable from hydraulic load 5. In addition, actual system pressure p.sub.ist can be measured with a pressure sensor 6 for example. Thus, in the course of controlling motor control unit 7 of electric motor 2, typically an inverter, control unit 4 specifies electrical motor torque M.sub.motor (or equivalently also a motor current), by means of which mechanical load torque M.sub.last, dependent on pump 3 or hydraulic load 5, sets in at electric motor 2.

[0024] The actual electric motor torque M.sub.motor can be estimated in a known manner by means of the motor current flowing through the windings of electric motor 2. However, mechanical load torque M.sub.last differs from electrical motor torque M.sub.motor, e.g., by an accelerated inertia of the motor plus friction losses, and is thus generally less than electrical motor torque M.sub.motor The mechanical load torque M.sub.last actually occurring between electric motor 2 and pump 3 is typically not measured in a servo drive 9 and can therefore also not be limited directly, which is why in prior art, fixed limits are provided at control unit 4 for electrical motor torque M.sub.motor. However, according to the invention a limiting unit 41 is provided, which uses an available, for example measured, dynamic system variable x of hydraulic system 1 to limit, based on that, motor torque M.sub.motor delivered by control unit 4 to servo drive 9, as shown by the arrow in FIG. 1. Since dynamic system variable x depends on the state of hydraulic system 1 and is thus directly or indirectly dependent on motor torque M.sub.motor and thus also on the occurring load torque M.sub.last, it is thereby possible to dynamically limit motor torque M.sub.motor. This limiting of motor torque M.sub.motor is thereby not rigidly specified as in prior art, but can vary from moment to moment as a function of the respective operating state of the hydraulic system. In this way, motor torques M.sub.motor can be applied to electric motor 2 of servo drive 9, which would not have been permitted in the case of rigidly specified limits of motor torque M.sub.motor. However, for these elevated or even negative motor torques M.sub.motor, system variable x must thereby result in no unpermitted state of hydraulic system 1, such as an excessively high or negative load torque M.sub.last.

[0025] As indicated in FIG. 1 by the dashed arrows, limiting unit 41 can obtain system variable x from any component of hydraulic system 1, for example control unit 4, servo drive 9, hydraulic load 5, etc. A redundant construction is also conceivableto the extent the system variable x of this component is available or can be derived from it.

[0026] FIG. 2 depicts an advantageous embodiment. An estimated load torque M.sub.last,ber, which approximates actual mechanical load torque M.sub.last, is determined in a calculation unit 42 via system variable x. Calculated mechanical load torque M.sub.last,ber represents an adequate estimate of the currently occurring actual mechanical load torque M.sub.last, by means of which dynamic load torque M.sub.last can be limited by limiting motor torque M.sub.motor. Calculation unit 42 thus serves as a type of shaft torque monitor.

[0027] Advantageously, motor speed .sub.motor can serve as system variable x; naturally, other or additional system variables x of hydraulic system 1 can be used to estimate mechanical load torque M.sub.last.sub._.sub.ber in limiting unit 41, for example volume flow V or electrical motor torque M.sub.motor, and so on.

[0028] In FIG. 2, there is also advantageously provided a comparison unit, which is connected to limiting unit 41 and calculation unit 42, and is advantageously an integral component of limiting unit 41. Comparison unit 43 is given a maximum load torque threshold M.sub.last,max and/or a minimum load torque threshold M.sub.last,min. Comparison unit 42 verifies whether estimated load torque M.sub.last exceeds or is likely to exceed maximum load torque threshold M.sub.last,max and/or a minimum load torque threshold M.sub.last,min, and in the event of a overshoot or a pending overshoot, it delivers a signal s to limiting unit 41. Limiting unit 41 then limits motor torque M.sub.motor, upon receiving signal s. This means that in this case even if control unit 4 specifies a higher or lower motor torque M.sub.motor, the limiting unit 41 does not forward these motor torque M.sub.last values to the servo drive 9, for example to prevent an impermissible load torque M.sub.last. As shown in FIG. 2, comparison unit 43 and calculation unit 42 may be an integral component of the limiting unit 41, or they may also represent independent components. Limiting unit 41 may also be an integral component of control unit 4 or motor control unit 7 of electric motor 2.

[0029] To implement the shaft torque monitor in limiting unit 41, servo drive 9 can be modeled as a control loop using the following model:

[00003]${\stackrel{.}{}}_{\mathrm{motor}}=\frac{k}{2.Math..Math..Math.J_{\mathrm{ges}}}.Math.{}_{\mathrm{motor}}+\frac{1}{J_{\mathrm{ges}}}.Math.M_{\mathrm{motor}}-\frac{1}{J_{\mathrm{ges}}}.Math.M_{\mathrm{last}}$

.sub.motor thereby refers to the motor speed, k is the torque constant, J.sub.ges is the known moment of inertia, M.sub.motor is the electrical motor torque and M.sub.last is the mechanical load torque. From this model, one can determine through conversion an estimated mechanical load torque M.sub.last as an approximation of mechanical load torque M.sub.last at the motor shaft of electric motor 2, or pump shaft of pump 3.

[00004]$M_{\mathrm{load}}=M_{\mathrm{motor}}-\frac{k}{2.Math..Math.}.Math.{}_{\mathrm{motor}}-J_{\mathrm{ges}}.Math.{\stackrel{.}{}}_{\mathrm{motor}}$

[0030] The mechanical load torque results from the motor torque decreased by a factor, which stems from a viscous friction of the pump and an acceleration of inertia.

[0031] Electric motor torque M.sub.motor as a calculated control variable is naturally known to the control unit 4, as is the motor speed the servo drive 9, which is normally provided by the servo drive 9 and serves as variable x. The moment of inertia J.sub.ges of the servo drive 9 includes the moment of inertia of the motor J.sub.motor, moment of inertia of the coupling J.sub.coupling (if present) and the moment of inertia of the shaft J.sub.shaft, which are known or can be drawn from data sheets of the respective components. The moment of inertia of the motor J.sub.motor thereby represents the dominant portion of the moment of inertia J.sub.ges, with which the inertial torque J.sub.ges is also approximated by the inertial motor torque J.sub.motor of the electric motor 2.

[0032] In actual practice, it has been found that the torque constant k.sub.0 specified by the manufacturer over the work range of the electric motor 2 deviates from the actual torque constant k. This also results in considerable inaccuracy when calculating the calculated load torque M.sub.last,ber. To reduce this inaccuracy, it may be provided to use a corrected torque constant k.sub.v instead of the specified torque constant k. To do so, one can proceed as follows.

[0033] To determine the transmission behavior of the drive line, i.e., of the electrical motor torque M.sub.motor on motor speed .sub.motor, an excitation signal can be applied to the drive line and one can measure the system response (motor speed) and from that, one can determine in a known manner a frequency response (as a Fourier-transform of the impulse response). In doing so, it was found that in servo pumps the amplitude response A1 of the frequency response corresponds approximately to known amplitude response A2 of a simple inertial mass with viscous friction, as shown in FIG. 3. This realization allows one to conclude that the transmission behavior of the drive line can be described by a simple inertial mass with viscous friction. With this knowledge, one can calculate a corrected torque constant k.sub.v, to better approximate actual torque constant k than would have been the case using specified torque constant k.sub.0.

[0034] To determine the corrected torque constant k.sub.v based on this knowledge, one can first represent the shaft output P.sub.shaft at the pump shaft of pump 3 as the product of torque M, factor 2 and rotation speed n in 1/minutes divided by 60:

[00005]$P_{\mathrm{weiie}}=\frac{M_{\mathrm{last}}.Math.2.Math..Math..Math.n}{60}.$

In contrast, the output P.sub.pump of pump 3 itself is calculated by the product of pressure p, pump volume per minute Q divided by 600 multiplied by pump efficiency .sub.pump:

[00006]$P_{\mathrm{pump}}.Math.\frac{p_{\mathrm{ist}}.Math.Q}{600.Math.{}_{\mathrm{pump}}}.$

If shaft output P.sub.welle and pump output P.sub.pump are made equal based on the conservation of energy, the equation

[00007]$\frac{M_{\mathrm{last}}.Math.2.Math..Math..Math.n}{60}=\frac{P_{\mathrm{ist}}.Math.Q}{600.Math.{}_{\mathrm{pump}}}$

results, which can be solved according to mechanical load torque M.sub.last. In this way, one obtains the mechanical load torque M.sub.last at an operating point from the product of the constant theoretical pumping volume of pump V.sub.th=Q/n, e.g., V.sub.th=160, 1 cm.sup.3/rev, and actual system pressure p.sub.ist, divided by pump efficiency .sub.pump multiplied by 20:

[00008]$M_{\mathrm{last}}=\frac{V_{\mathrm{th}}.Math.p_{\mathrm{ist}}}{20.Math..Math..Math.{}_{\mathrm{pump}}}.$

Pump efficiency .sub.pump can in turn be determined from the pump curve at the operating point, i.e., at a certain motor speed n. The pump curve represents a typical trend of the electrical motor torque M.sub.motor of the electrical motor 2 and the mechanical load torque M.sub.last of the electric motor 2, or pump 3, and is normally provided by the manufacturer of servo drive 9. In this way, given a rotation speed n=35 s.sup.1 as an operating point, a pump efficiency .sub.pump of 0.85 can be read.

[0035] Given an actual pressure p.sub.ist=139.1 bar, a rotation speed n=35 revolutions/s and a pump efficiency .sub.pump, of 0.85 (i.e., also a factor 10.85=0.15 in losses) results, i.e., in a torque decrease M.sub.v in the amount of 62.54 Nm, which is thereby proportional to the losses (1.sub.pump). Taking into account the viscous work, the corrected motor constant k.sub.v thus results in a value of k.sub.v=1.8 Nms when dividing torque decrease M.sub.v by rotation speed n=35 1/s:

[00009]$M_v=k_v.Math.n=\frac{\left(1-{}_{\mathrm{pump}}\right).Math.V_{\mathrm{th}}.Math.p_{\mathrm{ist}}}{20.Math..Math..Math.{}_{\mathrm{pump}}}.Math..Math.k_v=\frac{\left(1-{}_{\mathrm{pump}}\right).Math.V_{\mathrm{th}}.Math.p_{\mathrm{ist}}}{20.Math..Math..Math.{}_{\mathrm{pump}}.Math.n}=\frac{0.15.Math.160.1.Math.139.1}{20.Math..Math..Math.0.85.Math.35}.Math..Math.\mathrm{Nms}=\frac{62.54.Math..Math.\mathrm{Nms}}{35}=1.8.Math..Math.\mathrm{Nms}$

Corrected torque constant k.sub.v can also be used at the selected operating point for determining, according to the invention, the calculated mechanical load torque:

[00010]$M_{\mathrm{last},\mathrm{ber}}=M_{\mathrm{motor}}-\frac{k_v}{2.Math..Math.}.Math.{}_{\mathrm{motor}}-J_{\mathrm{motor}}.Math.{\stackrel{.}{}}_{\mathrm{motor}}$

[0036] Motor torque M.sub.motor is normally calculated from the product of a motor constant kt and a torque-forming current I. Motor constant kt can be optimized in a known manner. FIG. 4 contrasts conventional trends of calculated motor constants k.sub.t,calc, optimized motor constant K .sub.t,mod and actual measured motor constant k.sub.t,meas, wherein the respective values were standardized to amperes.

[0037] Using a corrected torque constant k.sub.v, FIG. 5 contrasts calculated mechanical load torque M.sub.last,ber,v and actual load torque M.sub.last for an operating cycle of a servo pump. Actual load torque M.sub.last was thereby reverse calculated from the measured system pressure .sub.ist. One can see that calculated mechanical load torque M.sub.last,ber,v using corrected torque constant k.sub.v, offers a better approximation of actual load torque M.sub.last than calculated mechanical load torque M.sub.last,ber which was calculated using specified torque constant k.sub.0.

[0038] FIG. 6 depicts a typical trend of an estimated load torque M.sub.last,ber. In FIG. 6, one can also see that motor torque M.sub.motor can experience peaks that would extend beyond a permitted load torque M.sub.last (for clarity's sake, not drawn in in FIG. 6), yet they do not result in any impermissible values of calculated load torque M.sub.last,ber and thus also of actual load torque M.sub.last. This is possible because, as mentioned, load torque M.sub.last responds in a delayed manner to motor torque M.sub.motor due to mass inertia, for example. Motor torque M.sub.motor is limited only when for example an impermissible estimated load torque M.sub.last,ber is reached, which is advantageous in relation to the fixed limits for motor torque M.sub.motor on control unit 4. An approximation or overshoot of the estimated load torque M.sub.last,ber in regard to the minimum and/or maximum load torque threshold M.sub.last,max, M.sub.last,min may be perceived as an impermissible estimated load torque M.sub.last,ber. As shown in FIG. 2 for example, estimated load torque M.sub.last,ber is calculated by calculation unit 42 using a system variable x (e.g., motor speed .sub.motor), compared by comparison unit 43 against the minimum or maximum load torque threshold M.sub.last,max, M.sub.last,min. In the event of an imminent overshoot, comparison unit 43 emits a signal s to limiting unit 41, which only then limits motor torque M.sub.motor. As soon as estimated load torque M.sub.last,ber begins to take on values again where no overshoot/undershoot of the minimum or maximum load torque thresholds M.sub.last,max, M.sub.last,min is imminent (comparison by comparison unit 43) based on the current value of the system variable in the course of the calculation by calculation unit 42, motor torque M.sub.motor is then no longer limited by limiting unit 41. Advantageously, this is signaled to limiting unit 41 by comparison unit 43 by the absence of signal s, another signal, or by other means.

[0039] It is hereby also possible in particular for the purpose of load shedding, in other words to quickly stop servo drive 9, to apply a negative motor torque M.sub.motor as long as the direction of rotation M.sub.last,ber does not change the sign. Before load torque M.sub.last becomes negative, the negative motor torque M.sub.motor is switched off.

[0040] Therefore, load torque M.sub.last monitoring according to the invention allows one to operate servo drive 9 in a more dynamic manner.