EP2275683B1 - Method for controlling a gear pump - Google Patents

Description

The present invention relates to a method according to the preamble of claim 1 and an application of the method.

Gear pumps consist of two intermeshing gears which are mounted on shafts, usually a shaft is connected to a drive unit. The shaft, which is not driven by a drive unit, is driven by means of torque transmission from the driven shaft via the tooth flanks.

Due to the torque transmission occurs due to excessive surface pressures often wear problems on the one hand, because on the tooth flanks of the driven shaft on the driven shaft, the pressure load caused by the differential pressure must be transmitted, and on the other hand, because the Friction work must be overcome. In particular, in the production of polymers in large polymerization or in the compounding of plastics with high throughputs at very high backpressures and high temperatures, ie overall high torque, damage to the tooth flanks by wear or abrasion (pittings, micro welding, wear) on the surface of the Gears occur.

In order to avoid this damage, two-shaft drives have already been used, in which the drive is effected by a single drive unit (motor), and then the power is then distributed by mechanical distribution gear to the two gear pump shafts.

Furthermore, from the patent CH-659 290 a gear pump known in which the two shafts are each driven by a drive unit. Each of the two gears receives the required drive power from the associated drive unit. Between the two gears only relatively minor difference performances are transmitted.

From the EP-0 886 068 B1 a gear pump is known, in turn, two drive units are provided for individually driving the waves, wherein phase and angular velocity of the intermeshing gears are coordinated such that on the one hand lifting of tooth flanks of intermeshing gears and on the other hand too high excess torque on the tooth flanks Interlocking gears are avoided.

It has been shown that in the known gear pumps, the wear, especially when conveying abrasive media, can be significant.

In particular, in extrusion applications of highly filled, abrasive polymer melts, the problem of high tooth flank wear by abrasion and thus the premature failure of the gear shafts because between the tooth flanks the abrasive particles contained in the melt are crushed and it can cause damage and erosion of the tooth flank surfaces. In addition, the viscosity of the melt and thus the torque requirement of the total pump or the torque required on the individual shafts increases as a result of filler loading, so that a possible overshoot of the permissible surface pressure on the tooth flanks again becomes the focus.

It is therefore an object of the present invention to provide a method of controlling a gear pump in which an improvement with respect to at least one of the mentioned disadvantages is obtained.

This object is achieved by the features stated in the characterizing part of claim 1. Advantageous embodiments and an application are specified in further claims.

The present invention initially relates to a method for controlling a gear pump consisting of two intermeshing gears, in which the two gears are each driven via corresponding shafts with a drive unit. The invention is characterized in that a momentary position of the one gear with respect to a current position of the other Gear is determined and that the current position of the one gear with respect to the current position of the other gear after setting predefined operating conditions continuously set and that the determination of the current position of the one gear with respect to the current position of the other gear set via a reference value which is determined prior to normal operation of the gear pump or during breaks in normal operation of the gear pump. Further embodiments of the present invention are characterized in that the reference value in the middle between tooth flanks of a tooth gap of a toothed wheel, preferably in the middle between tooth flanks of a tooth gap of a toothed wheel, is located.

• dass das eine Zahnrad das andere Zahnrad mit einem vorgegebenen Drehmoment antreibt,
• dass eine erste Winkeldifferenz durch Differenzbildung zwischen mit den Drehgeber/Sensoreinheiten gemessenen Werten bestimmt wird,
• dass das andere Zahnrad das eine Zahnrad mit dem vorgegebenen Drehmoment antreibt,
• dass eine zweite Winkeldifferenz durch Differenzbildung zwischen mit der Drehgeber/Sensoreinheiten gemessenen Werten bestimmt wird,
• dass eine Differenz zwischen der ersten Winkeldifferenz und der zweiten Winkeldifferenz bestimmt wird und
• dass der Referenzwert innerhalb der bestimmten Differenz festgelegt wird.

Further embodiments of the present invention are characterized in that the reference value is determined by

• in that one gear drives the other gear with a predetermined torque,
• a first angular difference is determined by subtraction between values measured by the encoder / sensor units,
• that the other gear drives the one gear with the predetermined torque,
• that a second angular difference by subtraction between the Encoder / sensor units measured values is determined
• a difference between the first angular difference and the second angular difference is determined, and
• that the reference value is set within the determined difference.

Thus, a method for automatic calibration of the arrangement is indicated with a gear pump. The system can carry out this calibration both before commissioning and during interruptions to operation without further action by the operating personnel. With this method, a possible wear of tooth flanks can be determined, because then resp. the difference between the first angular difference and the second angular difference also increases. An excessive occlusion can then be detected by a simple threshold violation.

• die erste Winkeldifferenz,
• ein Unterschied zwischen der ersten Winkeldifferenz und dem Referenzwert,
• die zweite Winkeldifferenz,
• ein Unterschied zwischen der zweiten Winkeldifferenz und dem Referenzwert,

und dass bei Unter- bzw. Überschreitung des mindestens einen der Momentanwerte unter einen vordefinierten Wert mindestens eine der folgenden Aktionen durchgeführt wird:

• eine optische Warnung,
• optische Anzeige,
• akustische Warnung,
• Änderung der Betriebsbedingungen der Zahnradpumpe.

Further embodiments of the present invention are characterized in that at least one of the following instantaneous values is monitored:

• the first angular difference,
• a difference between the first angular difference and the reference value,
• the second angular difference,
• a difference between the second angular difference and the reference value,

and that when the at least one of the instantaneous values is undershot or below a predefined value, at least one of the following actions is performed:

• an optical warning,
• optical display,
• acoustic warning,
• Change of the operating conditions of the gear pump.

Further embodiments of the present invention are characterized in that for determining the current positions of the one and the second gear shaft encoder / sensor units are used, each encoder / sensor unit are arranged centrally between the teeth of the respective gear and a rotor of the respective drive.

A central arrangement of the encoder / sensor units has the advantage that an existing angle of rotation due to a non-ideal rigidity of the entire drive train has a reduced influence on the measurement error of the system. The measurement error is halved by the central arrangement.

Further embodiments of the present invention are characterized in that a predefined flank play between two intermeshing gears is set.

Further embodiments of the present invention are characterized in that in the direction of rotation of one gear vorseilende flank of a tooth gap in a tooth gap touches a trailing in the direction of rotation of the other gear edge and that in the direction of rotation of the other leading edge of a tooth dipping into a tooth gap tooth touches a trailing in the direction of rotation of a gear edge.

This operating condition is also referred to as an edge change, because the touching tooth flanks change in the course of a Ausquetschvorganges.

Further embodiments of the present invention are characterized in that one gear drives the other gear with a predetermined torque, wherein the predetermined torque is greater than half of the total torque generated by the two drives.

An exact torque adjustment can be achieved by appropriate control of the speeds or instantaneous positions of the gears to each other. The tooth flanks thus transmit an arbitrarily adjustable torque, but never lift off each other during operation, whenever a defined flank seal is to be achieved. Further embodiments of the present invention are characterized in that the rotational speed of the shafts driven by the drive units is set so synchronously that a delivery medium pressure on a delivery side of the gear pump is substantially constant.

This has the advantage that no annoying pulsation on the pressure side of the gear pump is no longer present, which is reflected in the quality of the extrudate.

Further embodiments of the present invention are characterized in that the delivery medium pressure is measured on the pressure side of the gear pump and that the speed is adjusted in dependence on the measured delivery medium pressure.

Furthermore, the inventive method is characterized by the use of an arrangement with a gear pump comprising a pump housing, two housed in the pump housing and intermeshing gears and two shafts, which are operatively connected to the gears and are guided by the pump housing, wherein the two shafts are each operatively connected to a drive unit, wherein between the gear and the drive unit in each case a coupling unit for equalizing eccentricities between the drive unit and the respective shaft is arranged and wherein between the gear center and the drive center in each case a rotary encoder / sensor unit is arranged.

An embodiment variant of the present invention is characterized in that the rotary encoder / sensor unit lies in an axial region which is defined by the center between the center of the gear and the center of the drive plus a deviation of at most 10% of the distance between the center of the gear and the center of the drive.

Further embodiments of the present invention are characterized in that the rotary encoder / sensor units are each arranged in the middle between the respective gear center and the respective drive center.

Further embodiments of the present invention are characterized in that the rotary encoder / sensor unit to the axis of rotation of the respective shaft have a radial distance which is larger, preferably at least twice as large as an outer radius of the gears. Further embodiments of the present application are characterized in that the rotary encoder / sensor units are either optical or magnetic rotary encoder / sensor units. Other embodiments of the present application are characterized in that the rotary encoder / sensor units are arranged such that a perpendicular to the shaft and extending through the corresponding encoder / sensor unit connecting line with a centrally extending between the two axes of rotation suction side at an angle in the range of 35 ° to 55 °, preferably 40 ° to 50 °, preferably 45 °, including.

Further embodiments of the present application are characterized in that each drive unit comprises a rotor and a stator, wherein the rotor is axially displaceable with respect to the stator.

Further embodiments of the present application are characterized in that the drive units each have a compensating bearing unit on the side facing away from the gear pump, which radially supports the respective rotor of the drive unit.

Further embodiments of the present application are characterized in that the rotor of the drive unit is connected via the coupling unit with the respective shaft of the gear pump.

Further embodiments of the present application are characterized in that the coupling unit is a diaphragm coupling.

Further embodiments of the present application are characterized in that a flange between the pump housing and the stator of the respective drive unit is arranged, wherein the flange holes has, through which circulates a cooling medium for adjusting the temperature.

Further embodiments of the present application are characterized in that the drive units with the respective shaft of the gear pump is connected from the applied with respect to the gear pump side.

Further embodiments of the present application are characterized in that the connections between the drive units and the respective shafts of the gear pump are conical polygonal connections.

Further variants of the present application are characterized in that the drive units are of the torque motor type.

Further embodiments of the present application are characterized in that the one drive unit, the gear pump and the other drive unit are each contained in a temperature zone in which the temperatures are adjustable to predetermined values, wherein between adjacent temperature zones preferably isolation areas are present.

Fig. 1

eine bekannte Anordnung mit einer Zahnradpumpe und einer Antriebseinheit,

Fig. 2

einen Schnitt durch die in Fig. 4 angegebene Schnittebene A-A durch eine erfindungsgemässe Anordnung mit einer Zahnradpumpe und einer Antriebseinheit,

Fig. 3

eine schematische Darstellung der erfindungsgemässen Anordnung mit Angaben zu Temperaturzonen,

Fig. 4

eine Position für Drehgeber und Sensoreinheit zur Bestimmung von momentanen Positionen der Zahnräder,

Fig. 5

einen Schnitt quer durch die Drehachsen der Wellen im Bereich der Zahnräder zur Illustration einer ersten Betriebsbedingung,

Fig. 6

wiederum einen Schnitt quer durch die Drehachsen der Wellen im Bereich der Zahnräder zur Illustration einer zweiten Betriebsbedingung,

Fig. 7

wiederum einen Schnitt quer durch die Drehachsen der Wellen im Bereich der Zahnräder zur Illustration einer dritten Betriebsbedingung,

Fig. 8

wiederum einen Schnitt quer durch die Drehachsen der Wellen im Bereich der Zahnräder zur Illustration einer vierten Betriebsbedingung und

Fig. 9

einen Graphen mit einem Drehzahlverlauf, einem Druckverlauf und einem Drehmomentverlauf in Funktion der Zeit.

The invention is explained below with reference to drawings, in which embodiments of the present invention are illustrated. Showing:

Fig. 1

a known arrangement with a gear pump and a drive unit,

Fig. 2

a section through the in Fig. 4 indicated section plane AA by an inventive arrangement with a gear pump and a drive unit,

Fig. 3

a schematic representation of the inventive arrangement with information on temperature zones,

Fig. 4

a position for rotary encoder and sensor unit for determining instantaneous positions of the gears,

Fig. 5

a section across the axes of rotation of the waves in the region of the gears for illustrating a first operating condition,

Fig. 6

again a section across the axes of rotation of the shafts in the region of the gears to illustrate a second operating condition,

Fig. 7

again a section across the axes of rotation of the shafts in the region of the gears to illustrate a third operating condition,

Fig. 8

again a section across the axes of rotation of the waves in the region of the gears to illustrate a fourth operating condition and

Fig. 9

a graph with a speed curve, a pressure curve and a torque curve as a function of time.

In Fig. 1 a known arrangement is shown with a gear pump 1, the conveying medium F from a suction side S on a pressure side D promotes. It is evident in Fig. 1 a pump housing 10, are guided by the waves 2 and 3 to the outside. The outwardly guided shaft 3 is connected via a first universal joint 4, an adjustable in length axle 6 and a second universal joint 5 with a drive unit 7. Accordingly, the guided outward shaft 2 via a corresponding first and second universal joint and a corresponding axis section with a further drive unit (in Fig. 1 not shown). Thus, connected to the shafts 2, 3 gears (in Fig. 1 not visible) each driven by its own drive unit.

It should be noted that the double universal joint consisting of the first and second universal joint 4 and 5 is provided together with the adjustable axle section 6 for receiving lateral and angular deviations of the drive unit with respect to the shaft 2 and 3, respectively. Due to the double shaft joint in combination with the adjustable Achsabschnitt 6 acts an additional bearing force on a shaft 10 contained in the pump housing shaft bearing. This additional bearing force is due to the dead weight of the double-pivot joint and the axle section 6. The additional bearing force is due to a relatively short bearing distance of the pump bearings, which are in the pump housing 10 for supporting the shafts 2 and 3, with respect to the length of the double Swing joint considerably.
In Fig. 2 is a section through an inventive arrangement shown with a gear pump 1, wherein the cutting plane is placed in the axes of rotation 13 and 14 of the shafts 2 and 3 and by a sensor 25, according to the in Fig. 4 Plotted sectional plane AA. Fig. 2 shows the sake of simplicity, only one half of the gear pump 1. Accordingly, only one drive unit 7 is shown. The drive unit 7 is pressed directly via a flange 15, ie without intermediate gear, to the pump housing 10 or its lid. Via a screw 21, the rotating parts of the drive unit 7, such as a hub 16, a diaphragm coupling 22 and a rotor 18, connected to the shaft 3 of the gear pump 1. The screw 21 can be released if necessary, whereby the drive unit 7 can be solved by the gear pump 1 again. After loosening screws 40, which connects the flange 15 with the pump housing 10 or with its lid, and after loosening the screw 21, the complete drive unit 7 can be released from the gear pump 1. The waves 2,3 of the gear pump and their storage units stay within the gear pump and can be disassembled individually.

The drive unit 7 is in addition to the flange 15 and the hub 16 further comprises a rotor 18, a stator 17 and a drive cover 19 with an opening 20. The drive cover 19 closes the drive unit 7 on the side facing away from the gear pump 1 and is connected to the stator 17, wherein the opening 20 is arranged centrally on the extended axis of rotation 13 of the shaft 3. Gear pump side, the stator 17 is connected to the flange 15.

As has already been pointed out, the gear pump 1 is directly, ie without intermediate gear, connected to the drive unit 7. For this purpose, the screw 21 is provided by means of which the rotor 18 is fixed axially via the hub 16 and the flange 15. The screw 21 is guided during the assembly of the drive unit 7 to the gear pump 1 through the opening 20 in the drive cover 19 along the axis of rotation 13 of the shaft 3 and fixed in a corresponding bore in the shaft 3. In this case, the hub 16 is connected to the shaft 3 via a so-called conical polygon connection, on the one hand allows a precise axial alignment of the rotor 18 to the shaft 3, on the other hand, a very torsionally rigid connection between the rotor 18 of the drive unit 7 and the driven shaft 3 of the gear pump 1 allows.

It is already from the comparison of the known arrangement according to Fig. 1 and the inventive arrangement according to Fig. 2 clearly seen that the inventive arrangement is comparatively extremely short and due to the short Drehachsverbindung between the rotor 18 and the gear 11 also results in a very torsionally rigid connection, which is particularly in connection with the process of the invention to be explained is of importance.

As with the double shaft joint according to Fig. 1 also in the arrangement according to Fig. 2 an angular and a lateral compensation is required, on the one hand a diaphragm coupling 22 at the gear pump side end of the rotor 18 - for the angle compensation - and on the other hand, the stator 17 and the rotor 18 is formed such that the rotor 18 with respect to the stator 17 axially is slidable to allow the lateral compensation.

The membrane coupling 22 and the hub 16 are conceivable, for example, as a single part, as well as from Fig. 2 shows that in the left, drive side half of the individual part of the classic function of a hub which can be coupled to the shaft 3, in the right part of this single part is thin-walled and thus fulfills the functions of a membrane coupling.

In addition to the mentioned support of the rotor 18 with respect to the stator 17 on the side of the gear pump 1 by means of flange 15 and hub 16 - via conical polygonal connection and screw 21 - is on the with respect to the gear pump. 1 side facing away from a compensating bearing 23 is provided, which holds the rotor 18 in relation to the stator 17 radially in position.

Production-related eccentricities of the rotational axis 13 of the shaft 3 with respect to an axis of rotation of the compensating bearing 23 can be compensated by the diaphragm coupling 22. Although due to production-related eccentricities an additional burden on the gear pump bearings, but the resulting torque reactions hold within relatively narrow limits, since the distance of the diaphragm coupling 22 to the loaded bearing is low and since only a moderate angle compensation must be accomplished.

As a drive unit 7, for example, a so-called torque motor is used, which is a high-pole permanent-magnet three-phase synchronous motor with hollow shaft rotor for the direct, above-mentioned coupling to the gear pump. Torque motors are characterized in particular by a short compact design and a low torsional backlash (high torsional rigidity).

As will become apparent from the following comments on the operation of the arrangement with the gear pump 1, precise information about the current position of the one gear in relation to the current position of the other gear are very important. Along with this is the demand for a direct and possibly undistorted influence of the drive units on the gears of the gear pump. One The criterion is the already mentioned low torsional backlash (high rigidity) between the drive unit and the driven gear. Another criterion is the most accurate possible measurement of the current position of the one gear with respect to the current position of the other gear.

In the embodiment according to Fig. 2 This is achieved in that at the periphery of the hub 16, a rotary encoder 24 is arranged, which cooperates with a connected to the stator 17 sensor unit 25. For example, a grid is applied to the hub 16, which is read by the sensor unit 25. Instead of such an optical measuring device, corresponding magnetic measuring devices or other methods for position determination can also be used.

In order to minimize any measurement errors due to eccentricity of the rotary encoder 24 to the toothing, the rotary encoder 24 is made as large as possible in diameter. The eccentricity of the encoder 24 itself is minimized by the integration of the inclusion of the encoder 24 in the hub 16. Since the hub 16 is in one piece, very close manufacturing tolerances can be adhered to the inclusion of the encoder 24.

The position of the encoder 24 resp. The sensor unit 25 is preferably selected between the center of the rotor 18 or stator 17 and the center of the driven gear 11 of the gear pump 1. In a uniform Stiffness distribution over the drive train (ie between the center of the rotor 18 and stator 17 and the center of the driven gear 11 of the gear pump 1) is the encoder 24, respectively. the sensor unit 25 is preferably arranged in the middle between the center of the rotor 18 and stator 17 and the center of the driven gear 11 of the gear pump 1.

A possible field of application of the arrangement with a gear pump is the pressure build-up downstream of an extruder in the conveyance of plastic melts in an extrusion line. In this field of application, the polymer melts are conveyed at temperatures of up to 300 ° C. against high discharge pressures (eg 300 bar). For this purpose, high drive power and thus high torques are necessary. Accordingly, the gear pump or the pump housing is heated to a temperature of, for example, 300 ° C, caused by the conveying medium, while the temperature of the drive units 7 and 8, especially for the necessary electronic circuits in these, should not exceed 60 ° C. To illustrate this facts shows Fig. 3 in a schematic representation of the inventive arrangement with a gear pump 1, wherein now the gear pump 1 and the two laterally arranged drive units 7 and 8 are represented by simple blocks. The individual components are contained in temperature zones 32, 33 and 34, which according to the above explanations must have permitted or required temperature values. Thus, in the temperature zone 33, the gear pump 1 contained, which is operated due to the temperature of the pumped medium, for example at 300 ° C. On the drive side, the drive units 7 and 8 are provided, which in the temperature zones 32, respectively. 34 whose maximum value must not exceed 60 ° C for proper functioning. The present arrangement requires the placement of electrical components in the immediate vicinity of the gear pump. Since the gear pump is up to 300 ° C hot, insulating partitions 30 and 31 are required, which between the temperature zones 32 and 33, respectively. between the temperature zones 33 and 34 are present. In addition to the insulating partitions 30 and 31, additional measures are required as needed, so that the temperature in the cold temperature zones 32 and 34 does not reach unacceptable levels. An additional measure, for example, is that an active cooling (for example, an active water cooling) is provided.

It is also conceivable, the rotor 18 ( Fig. 2 ) to protect from over temperature by a cooling of the flange 15 is connected between the hub 16 and the gear pump 1. The cooling is realized for example by star-shaped holes in the flange 15. This achieves very good cooling properties, since the deflections produce high turbulences. The hub 16 is cooled on the entire surface flange side by radiation and forced convection.

Fig. 4 shows a possible positioning of the sensor unit 25, which is used to determine the current position of the one Gear is used in relation to the current position of the other gear, wherein Fig. 4 a section transverse to the axes of rotation 11 and 13 of the shafts 2 and 3 shows. The fluid is transported in the direction of arrow from the suction side S with the gear pump on the pressure side D. In this case, a force component is generated in the direction of the arrows P, P ', which act on the shaft bearing of the gear pump and to a slight displacement of the shafts 2 and 3 (FIGS. Fig. 2 ) to lead.

To compensate for the eccentricity caused by the displacement, the sensor unit 25 will now move in the direction of displacement, i. in the direction of deflection of the shaft, attached. The attachment takes place, for example, below 45 ° and is thus on average of the possible displacement angle, which is differential pressure dependent and viscosity-dependent. With insufficient accuracy, caused by the shaft displacement and deflection, z. B. a zahnradbreiten- and game size-dependent arrangement of the sensor unit 25 are made.

Based on Fig. 5 to 9 In the following, various operating conditions are explained, which can be specified as predefined sequences for the operation of the arrangement with the gear pump.

Fig. 5 shows a section transverse to the axes of rotation 13 and 14 in the region of the gears 11 and 12. delivery medium F is received on the suction side S of the tooth gaps and then along the pump housing to the pressure side D transported where the medium F is squeezed out by the meshing gears 11, 12.

During operation of the gear pump, a "trapped volume" is created in the toothed area between the tooth base and the tooth tip of the gears, which is sealed off by the almost touching tooth flanks in front of and behind this volume. In terms of flow technology, however, a flow gap can be generated selectively at the points where a large flow gap is desired for tribological reasons (optimum gap thickness relative to the relative speed of the tooth flanks). Due to the existing position control of the shafts, the ratio of these two sealing gaps can be actively controlled. Once, the gap leading to the "trapped volume" can be minimized, once the gap following the trapped volume. This makes it possible for the squeezing process to be actively influenced from this "trapped volume", thus optimizing the uniformity of the flow.

• In einem ersten Schritt treibt die erste Welle 2 die zweite Welle 3 mit einem definierten Drehmoment an. Dabei wird eine erste absolute Drehwinkeldifferenz mit Hilfe des erläuterten Drehgebers 24 in Kombination mit der Sensoreinheit 25 (Fig. 2) bei beiden Wellen 2 und 3 bestimmt, indem eine Differenz zwischen einem gemessenen Wert der einen Sensoreinheit 25 und einem gemessenen Wert der anderen Sensoreinheiten 25' bestimmt wird.

To the various operating conditions or
To be able to set operating conditions of the arrangement according to the invention, information about the current position of the one gear 11 with respect to the current position of the other gear 12 must be known. These details are the actual initial conditions that are necessary for further settings of the gears to each other. One way to determine this information is in the implementation of the following steps of the method, which is also referred to as calibration:

• In a first step, the first shaft 2 drives the second shaft 3 with a defined torque. In this case, a first absolute rotation angle difference with the aid of the described rotary encoder 24 in combination with the sensor unit 25 (FIG. Fig. 2 ) at both shafts 2 and 3 is determined by determining a difference between a measured value of one sensor unit 25 and a measured value of the other sensor units 25 '.

In a second step, the second shaft 3 drives the first shaft 2 with the same defined torque as in the first step. In this case, a second absolute rotation angle difference is in turn determined by means of the described rotary encoder 24 in combination with the sensor unit 25 in both shafts 2 and 3, again determining the difference between a measured value of the one sensor unit 25 and a measured value of the other sensor units 25 ' ,

In a third step, a difference between the first absolute rotation difference and the second absolute rotation difference is formed. This difference is the actual range in which the gears can move to each other, provided that the defined torque that was used in the first and second step in the measurement is not exceeded. In this area, a reference value can now be defined, in relation to which the current positions of the gears are specified. The reference value is then a zero point of a defined coordinate system. For example, the reference value lies in the middle between tooth flanks of a tooth gap, so that the absolute values of the maximum deflections are identical.

When using the Fig. 5 a constant force FO between the gears 11 and 12 is transmitted, as shown in the detailed representation X in the right half of Fig. 5 is illustrated.

Thus, after setting the reference point, the operating conditions can now be selected in a first setting, for example, so that one gear transmits half plus a defined percentage of the total torque. Accordingly, then transmits the other gear half minus the defined percentage of total torque.

Under these operating conditions, a defined seal between the tooth flanks can be achieved. The scope of these operating conditions is aimed at the promotion of low-viscosity fluids, in which a sealing effect between the gear edges is necessary in order to obtain a sufficient seal from the pressure side D to the suction side S.
Another setting is that as operating conditions, the backlash between the flanks of two intermeshing teeth is selectable, namely for example, in 10% increments from the contact of the flanks (no backlash) over a central alignment (ie, the tooth entering a tooth gap lies exactly in the middle of the gap) until the tooth flanks touch again, this time around the trailing tooth flanks is.

Fig. 6 illustrates the operating conditions just explained in turn in a section transverse to the axes of rotation 13 and 14 in the region of the gears 11 and 12. Again, in the meshing region of the gears 11 and 12, a detail X shown enlarged as detail, in which also an adjusted backlash 26th is shown highlighted.

These operating conditions are selected when the medium F has an average viscosity. With the adjustment of the backlash 26, the squeezing pressure can be adjusted so that it corresponds as equal as possible to the pressure on the pressure side D. Too large a backlash 26, which leads to a smaller squish pressure than the pressure on the pressure side D, must be avoided because too little sealing effect between the pressure side D and suction side S is obtained. The operating condition, in which a certain backlash 26 (ie without flank contact) is present, can then be realized with a corrosion-resistant (and thus often soft) coating of the gears, without damage caused by abrasion.

Another setting is that an edge change mode is proposed as operating conditions. In this case, a gear changes the flanks during the theoretical rolling of a tooth on the line of action. The Quetschdruckentladung thus always targeted to the suction side.

The edge change operation is based on Fig. 7 explained, in turn, sections transverse to the axes of rotation 13, 14 in a region of the meshing gears 11, 12 show. On the left half of Fig. 7 a state is shown in which the tooth Z ₁ 'of the toothed wheel 11, which engages in a tooth gap of the toothed wheel 12, contacts the tooth Z ₁ . On the right half of the Fig. 7 Then, a state shown later in time, in which the engaging in a tooth gap of the gear 11 tooth Z _{2 of} the gear 12, the tooth Z ₁ 'touches. Thus, the mentioned edge change has been made in the meantime.

• Minimierung des Quetschdruck-bedingten Pulsierens durch Quetschdruckentladung zur Saugseite S;
• Minimierung des aufzuwendenden Drehmomentes durch Minimierung der Quetschdruckenergie;
• Reduzierung der Temperaturerhöhung durch Minimierung der Quetschdruckenergie.

The edge change operation has at least one of the following advantages:

• Minimization of the squeezing pressure-related pulsation by squeezing pressure discharge to the suction side S;
• Minimizing the torque required by minimizing the squish pressure energy;
• Reduction of the temperature increase by minimizing the squish pressure energy.

The operating conditions according to the mentioned edge change operation are used, for example, in highly viscous media in which the squeezing pressure is so great that a very large torque is required to produce the squeeze pressure energy, as these represent a pure energy loss.

With the flexibility of electronic control, depending on the properties of the fluid being conveyed, i. the flow behavior or the solids loading of the polymer to be delivered, the Ausquetschverhalten be varied specifically. So each polymer type can be assigned an optimal speed profile.

Based Fig. 8 a further aspect of the method according to the invention is explained. Based on the knowledge of the current position of the one gear 11 with respect to the current position of the other gear 12, for example by applying the explained steps for a calibration, and the maximum margin (difference between the first absolute rotation difference and the second absolute rotation difference) for a momentary position of the one gear 11 with respect to the current position of the other gear 12, now a statement about a wear of the gears 11, 12 can be made, for example, the maximum margin at a certain, from a gear 11, 12th changed to the other gear 12, 11 transmitted torque. For example, if the margin over a given maximum threshold also increased, this can be interpreted so that soon a gear and / or shaft must be replaced with gear because, for example, with a speedy failure of the system must be expected. So are in a gear pump, according to the left half of Fig. 8 is adjusted so that always squeezed onto the suction side S, when the maximum allowable Verschleisswertes (ie the maximum allowable margin on one side, starting from the reference value) is exceeded, the operating conditions automatically or after appropriate manual confirmation of a supervisor switched. The conversion of the operating conditions can be carried out such that now the necessary seal is transmitted to the other tooth flanks. According to the right half of Fig. 8 If this is the case, the leading edge of the tooth is an anticipatory tooth flank.

It is also conceivable that upon detection of a wear, the instantaneous position of the one gear is changed with respect to the current position of the other gear so that the originally desired optimum operating conditions are maintained. For example, a lifting of the tooth flanks may occur due to wear. The corresponding correction to restore the desired operating conditions would then be a change in the current position of the one gear in relation to the current position of the other gear, so that the tooth flanks touch again in the desired manner or in turn the desired backlash is obtained.

The monitoring of signs of wear can also be exploited to the extent that upon detection of a predetermined level of wear an audible and / or visual warning is given to the supervisor, so that precautions can be taken to prevent failure of the pump system. Thus, it is conceivable that when issuing an appropriate warning, a replacement shaft or a spare gear is given to the manufacturer in time to order so that the required spare parts are available before a possible failure of the pump system on site.

In some extrusion plants in which gear pumps are used, pressure fluctuations due to the aforementioned Ausquetschvorganges of residual volume between the gears. These pressure fluctuations are also referred to as pulsations and lead to irregularities in the product produced by the extrusion. For this reason, various measures have already been proposed to reduce the pressure fluctuations. Worth mentioning is the use of helical gears or Pfeilverzahnungen, both of which, however, have system-related disadvantages.

According to the present invention, pressure fluctuations are eliminated or at least greatly reduced by actively influencing the rotational speeds of the two gear shafts.

The inventive arrangement or the inventive method is capable of the speed curve per Ausquetschvorgang to vary, in such a way that the pressure on the pressure side is within narrow limits resp. that the pressure on the pressure side is constant. Thus, the Ausquetschvorgang of the pumped medium from the tooth base is controlled specifically on the current position of the one gear with respect to the current position of the other gear.

It has been found that, although it is in principle desirable that the pressure fluctuations can be completely eliminated. However, in certain applications, pressure fluctuations may be desired specifically to provide appropriate variations in extrudate thickness. Thus, the inventive method, in particular in connection with the inventive arrangement with a gear pump, opens up new production possibilities in the extrusion.

In order to be able to obtain as simple as possible and at the same time complete elimination of pressure fluctuations, a degree of coverage of 1 must be selected. If an overlap degree of 1 is selected, then only one tooth pair is involved in the displacement, ie the squeezing out (cf. Vogel Fachbuch Jarosla and Monika Ivantysyn: "Hydrostatic Pumps and Motors", 1993, p. 319 ). In this case, a sinusoidal course of the displacement volume flow results. This can be easily and efficiently corrected via a sinusoidal compensation table (for example, a so-called "look-up" table).

In Fig. 9 a speed curve 90 of the gear pump shaft, a pressure curve 91 of the pressure on the pressure side of the gear pump and a torque curve 92 of the torque of the gear pump shaft is shown. The speed curve 90, the pressure curve 91 and the torque curve 92 are plotted as a function of time t. The speed of the gear pump shaft is adjusted as a function of time so that the pressure on the pressure side of the gear pump is constant or at least within a predetermined tolerance range. The in Fig. 9 shown speed curve 90 has a periodicity with a period T on. It is the period of time during which a tooth engagement takes place in a corresponding tooth gap. If, therefore, the speed for both shafts is controlled synchronously in accordance with the speed curve 90, the pulsation can be completely compensated.

It is expressly pointed out that the pulsation compensation can be combined with all operating conditions or specifications explained in this description.

Due to the periodicity, it is possible to store the speed curve 90 in a memory unit (Look-up Table). The values for the speed to be set are then read out in a predetermined cycle, the predetermined clock resulting from the pressure to be set on the pressure side.

Alternatively, it is also possible to measure the pressure on the pressure side with a pressure sensor, and to use the speed due to the measured pressure continuously to adjust the speed. Although this process, which is referred to as an on-line pressure adjustment method, is more complicated to implement, this results in further possible applications for realizing specific production processes in extrusion.

The selective influencing of the position control, as has been explained above for preventing or reducing the pulsation of the pressure on the pressure side of a gear pump, can also be used for shear-sensitive materials to reduce a total shear stress. Thus, when determining the speed curve, care is taken to ensure that a maximum permissible shear load is not exceeded.

The present invention makes it possible for the first time to specifically influence the effects of pulsation, crushing pressures and tribological behavior. The settings may take into account all effects that are relevant to the specific case, or individual operating conditions may be considered as a priority. By this is meant that these operating conditions should have a more significant influence on the behavior of the overall system.

The advantage of the involute toothings typically used for gear pumps is that the transmission ratio of the two rotational speeds remains constant during one revolution, which is a basic prerequisite for a constant volume flow. Circular-toothed gears, however, have the disadvantage that the transmission ratio of the rotational speeds of the waves fluctuates periodically and thus the delivery medium flow pulses. The use of the described invention with two controlled drive units makes it possible for the first time to use circular-arc toothing, without causing an undesired pulsation of the delivery medium flow. Thus, with sufficiently large backlash the drive speeds of the waves can be corrected accordingly and compensated with opposite speed profile, so that circular arcs with constant transmission ratio and thus constant flow are possible. As with the mentioned circular arc toothing other tooth shapes are conceivable. It is only necessary to adjust the speed profile accordingly.

Claims (14)

1. Method for controlling a gear pump (1) comprising two meshing gear wheels (11, 12), at which gear pump (1) the two gear wheels (11, 12) each are driven by a drive unit (7, 8) via corresponding shafts (2, 3), wherein a momentary position of the one gear wheel (11, 12) is determined with respect to a momentary position of the other gear wheel (12, 11), that the momentary position of the one gear wheel (11, 12) is continuously adjusted with respect to the momentary position of the other gear wheel (12, 11) according to predefined operating conditions, characterized in that the determination of the momentary position of the one gear wheel (11, 12) is adjusted with respect to the momentary position of the other gear wheel (12, 11) by means of a reference value, that is determined before the normal operation of the gear pump (1) or during interruptions of the normal operation of the gear pump (1).
2. Method according to claim 1, characterized in that the reference value lies between tooth flanks of a tooth space of a gear wheel (11, 12), preferably in the middle between tooth flanks of a tooth space of a gear wheel (11, 12).
3. Method according to claim 2, characterized in that the reference value is determined in that

   - the one gear wheel (11) drives the other gear wheel (12) with a specified torque,

   - a first angle difference is determined by differentiation between values measured by the rotary encoder/sensor units (24, 25),

   - the other gear wheel (12) drives the one gear wheel (11) with the specified torque,

   - a second angle difference is determined by differentiation between values measured by the rotary encoder/sensor units (24, 25),

   - the difference between the first angle difference and the second angle difference is determined and

   - the reference value is set within the determined difference.
4. Method according to claim 3, characterized in that at least one of the following momentary values is monitored:

   - the first angle difference,

   - the difference between the first angle difference and the reference value,

   - the second angle difference,

   - the difference between the second angle difference and the reference value,

   and in that when at least one of the momentary values is below resp. above a predefined value, at least one of the following actions is carried out:

   - an optical warning,

   - an optical display,

   - an acoustic warning,

   - change of the operating conditions of the gear pump (1).
5. Method according to one of claims 1 to 4, characterized in that rotary encoder/sensor units (24, 25) are used for the determination of the momentary position of the one and the other gear wheel (11, 12), wherein each rotary encoder/sensor unit (24, 25) is arranged in the middle between the toothing of the respective gear wheel (11, 12) and a rotor (18) of the respective drive unit (7, 8).
6. Method according to one of claims 1 to 5, characterized in that a predefined flank clearance (26) is adjusted between two meshing gear wheels (11, 12).
7. Method according to one of claims 1 to 5, characterized in that a leading flank with respect to the rotating direction of the one gear wheel (11) of a tooth immersing in a tooth gap touches a following flank with respect to the rotating direction of the other gear wheel (12) and in that a leading flank with respect to the rotating direction of the other gear wheel (12) of a tooth immersing in a tooth gap touches a following flank with respect to the rotating direction of the one gear wheel (11).
8. Method according to one of claims 1 to 5, characterized in that the one gear wheel (11, 12) drives the other gear wheel (12, 11) with a specified torque, wherein the specified torque is greater than half of the entire torque that is produced by the two drive units (7, 8).
9. Method according to one of claims 1 to 8, characterized in that the number of revolutions of the shafts (2, 3) driven by the drive units (7, 8) is adjusted synchronously as such that a pressure of the medium to be pumped is in essence constant on a pressure side (D) of the gear pump (1).
10. Method according to claim 8, characterized in that the pressure of the medium to be pumped is measured on the pressure side (D) of the gear pump (1) und in that the number of revolutions is adjusted in dependence of the measured pressure of the medium to be pumped.
11. Method according to one of the claims 1 to 10, characterized in that an arrangement with a gear pump (1) is used comprising a pump housing (10), two meshing gear wheels (11, 12) comprised within the pump housing (10) and two shafts (2, 3) that are operationally connected to the gear wheels (11, 12) and are guided through the pump housing (10), wherein each of the two shafts (2, 3) are operationally connected to a respective drive unit (7, 8), wherein a coupling unit (22) is arranged between each gear wheel (11, 12) and drive unit (7, 8) for the equalization of eccentricities between each drive unit (7, 8) and corresponding shaft (2, 3) and wherein a rotary encoder/sensor unit (24, 25) is arranged between each gear wheel centre and drive centre.
12. Method according to claim 11, characterized in that the rotary encoder/sensor unit (24, 25) is arranged in an axial region that is defined by the middle between the gear wheel centre and the drive centre plus a deviation on both sides of maximal 10% of the distance between the gear wheel centre and the drive centre.
13. Method according to claim 12, characterized in that the rotary encoder/sensor units (24, 25) each are arranged in the middle between each gear wheel centre and each drive centre.
14. Method according to one of claims 11 to 13, characterized in that rotary encoder/sensor unit (24, 25) comprises a radial distance to the axis of rotation (13, 14) of each shaft (2, 3), which is greater than an outer radius of the gear wheels (11, 12), preferably at least twice as great.

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