US20100143155A1

US20100143155A1 - Piston Pump Having A Force Sensor And A Method For Controlling Said Pump

Info

Publication number: US20100143155A1
Application number: US12/631,695
Authority: US
Inventors: Thomas PREISWERK; Rene CHAVAN
Original assignee: FLUX INSTRUMENTS AG
Current assignee: FLUX INSTRUMENTS AG
Priority date: 2008-04-02
Filing date: 2009-12-04
Publication date: 2010-06-10
Also published as: CN101865108A; EP2107241A2; EP2107241A3

Abstract

The present invention relates to a piston pump for generating a delivery flow, which is substantially free of pulsation, in particular to a dual piston pump, and to a method for controlling such a piston pump delivering a pumped medium from a low-pressure area into a high-pressure area, wherein measuring sensors detecting mechanical forces or moments transmitted by the structure of the pump or its associated drive unit are used instead of the pressure or flow sensors usually employed for this purpose.

Description

FIELD OF THE INVENTION

The present invention relates to a piston pump for generating a delivery flow which is substantially free of pulsation, in particular to a dual-piston pump, and to a method for controlling such a piston pump delivering a pumped medium from a low-pressure area into a high-pressure area.

BACKGROUND OF THE INVENTION

A piston pump according to the present invention comprises at least two piston/cylinder units for delivering the pumped medium from a low-pressure area into a high-pressure area, a cam drive for driving at least one piston/cylinder unit, a control unit for controlling the rotational speed of the cam drive, and a sensor for measuring an actual value of a control parameter, by which actual value the extent of pulsation of the delivery flow generated on the high-pressure side can be derived.
Pumps of this kind are employed, for example, in liquid chromatography, especially in high-pressure liquid chromatography (HPLC) and in ultrahigh-pressure liquid chromatography (UHPLC), for the delivery of the mobile phase (the eluant), for example, in form of a pure solvent or a low pressure side gradient solvent mixture through the stationary phase (package) in a separation column of a pertinent analysis system. The term “low pressure side gradient” will be familiar to each skilled person.

SUMMARY

Correct analysis results with regard to quality (retention time) and quantity (peak area) require a continuously constant mass flow of the mobile phase. Ideally, true mass flow should be ensured, i.e. a delivery flow, which is constant over a certain unit of time with respect to volume at atmospheric pressure. A true mass flow should be ensured at least for the duration of a series of analysis runs associated with each other. A constant mass flow is of critical significance, especially in HPLC and UHPLC, since the delivery pressures of said pumps may reach 100 MPa and more when using said analysis methods.
Under such operating conditions, the pumped medium no longer behaves as an ideal, i.e. incompressible liquid. This results in the fact that due to the specific compressibility of the pumped medium a rising delivery pressure increasingly exacerbates the generation of true mass flow of the pumped medium, which is substantially free of pulsation.
In order to reduce pulsations in a high-pressure delivery flow generated by a piston pump, for example, pumps having a plurality of, in particular two, pistons are employed operating in accordance with the parallel or preferably the serial delivery principle. In other words, the pistons deliver into a common high-pressure area, wherein their stroke movements are offset in time with respect to each other such that the individual delivery flows are theoretically superimposed over each other to form a composite constant flow, which is free of pulsation.
With low delivery pressures a pulsation of the total delivery flow can thus be completely or nearly avoided, with high and highest delivery pressures, however, a residual pulsation arises. This residual pulsation is caused by the specific compressibility of the pumped medium, which has an increasing effect as the delivery pressure rises and its extent depends on the changing physical properties of the pumped medium. Said residual pulsation may be reduced or suppressed by commencing each delivery cycle with a precompression stroke, which ensures that the medium within the cylinder internal space is compressed to match the respective system pressure prevailing on the high-pressure side prior to the onset of the actual delivery in order to avoid or at least minimize a short-term backflow of pumped medium and/or an interruption of the continuous delivery upon hydraulic transition from one piston to the other.
Feedback controls, known from pertinent pumps, for (continuously) compensating the influence of the specific compressibility of the pumped medium on the momentary pumping efficiency are either based on monitoring the delivery pressure or determining the flow rate (e.g. by sensors operating according to the caloric measuring principle) at the inlet and/or outlet of a pump as well as the qualitative or quantitative processing of the respective measuring signals.
U.S. Pat. No. 4,359,312 discloses a dual-piston pump having a cam drive operating in accordance with the parallel or serial delivery principle of the kind described above as well as a method for controlling said pump based on a subtractive approach. In order to be able to counteract a drop in the pumping efficiency at increased delivery pressures, the drive cam(s) causing the actual pumping action comprises an elevation profile section for generating a pre-compression stroke, with a suction stroke correspondingly shortened in relation to the angular range and accelerated due to construction. Initially, the length of the pre-compression stroke is chosen such that it compensates for the influence of the specific compressibility of the pumped medium on the pumping efficiency at a specified maximum specific compressibility and a specified maximum allowable delivery pressure the pump is driven so as to fully compensate for these maximum conditions. For this purpose the device comprises a pressure sensor by means of which the system pressure on the high-pressure side is continuously monitored.
The measured values monitored by the pressure sensor are fed into the drive control system and the rotational speed of the drive motor is modulated in the pre-compression stroke range such that, with an iterative and subtractive approach, the pump is iteratively backed off from compensation of the maximum conditions so as to compensate for the actual conditions of the pumped medium at the actual operating pressure. In parallel therewith, an increased delivery with respect to volume at atmospheric pressure is obtained due to the pre-compression stroke used respectively in the feedback process. This surplus delivery is corrected by means of readjustment by superimposing a secondary correction factor to the cam rotational speed.
In U.S. Pat. No. 4,681,513 another pump and another method are disclosed. Herein, each of the two pistons of a dual-piston pump operating in accordance with the serial delivery principle has a separate cam drive. Each of the two drive cams, in sections, has different elevation profiles characterizing different stroke lengths, which are graded according to certain maximum delivery rates of the pump, with their slopes exhibiting a constant rise. During each pumping cycle, within the associated angular range of the cams, both, the suction stroke and the pumping stroke as well as a preceding pre-compression stroke are caused by a reciprocating movement of the drive cams and the two delivery flows are combined such that a composite constant delivery is achieved.
A crucial inherent disadvantage of all feedback controls for compensation of compressibility (pressure feedback), which are based on a qualitative and/or quantitative utilization of the measurement of the delivery pressure, is the need for an iterative adjustment (PID control) within a series of cam revolutions/pumping cycles and a practically limited robustness of the control upon occurrence of pressure artefacts, since only an indirect and no direct monitoring of the process of displacement of the pumped medium is possible by the pressure measuring signal. Disturbance variables are, for example, pressure pulses, occurring when a sample is fed into the pertinent analysis system, and in particular the drift of the pressure measuring signal during gradient operation, wherein back-pressure changes arise or may arise at the separation column due to varying physical properties (viscosity; temperature) and/or when a clogging of the column occurs or may occur.
Apart from the pressure feedback method for compensating the influence of the specific compressibility of the actually pumped medium on the pumping efficiency, a flow feedback method is known, which is based upon monitoring the delivery characteristics of the pump by means of flow sensors arranged on the high-pressure side (and optionally also on the low-pressure side) within the liquid duct path and wherein the feedback method relies on measuring signals of said sensors.
Specific technical problems arise with the pressure measuring sensors as well as with the flow sensors from the fact that they must be suited to a very wide measuring range of 0-100 MPa for example and thus, must be able to withstand the high pressure and be peripherally sealable with acceptable effort. At the same time, they must not affect optimal configuration of the geometry of the liquid paths in terms of pressure drop and fluid dynamics and, since they are in direct contact with the medium to be delivered, they must either be chemically resistant to all media to be delivered or be protected from said media by a separation membrane or the like. In the case of flow sensors it is of particular disadvantage that the measuring signal is dependent upon the operating temperature as well as the specific thermal conductivity/capacity of the respective medium to be delivered and that consequently, a multiparametric calibration with regard to application is necessary, which can be achieved during operation of the pump in accordance with the low-pressure gradient method at best by a detour via the method of a learning-in run by which a set of reference values is assessed under real analysis run conditions.
The difficulties described above with regard to design concept superimpose the basic systematic problem of the decrease of the pumping efficiency by detrimental dead volume within the displacement system (piston/cylinder unit) of the working piston when considering the case of a serial dual-piston pump. Said volume being that which is not displaced from the displacement chamber—calculated (ranging) from the closing edge of the inlet valve to the closing edge of the outlet valve—during the pumping stroke and which, remaining therein, acts as a hydraulically elastic element, since liquids have a noticeable specific compressibility. Implementing the concept of all-in/all-out, which theoretically presents a solution to the problem, has technical limitations as to the respective design of the components belonging to the displacement chamber feasible in practice.
With increasing delivery pressure, the compressibility of the medium to be delivered causes an increasing drop of the pumping efficiency and associated with that a delivery flow having a more or less pronounced residual pulsation. The pumping efficiency is diminished solely as a function of the detrimental dead volume, which is technically difficult to minimize, and not as a result of the compression of the (actual) pumping volume during the displacement stroke.
When using the pump while applying the low-pressure gradient methodology, the detrimental dead volume also affects the proportionating of the individual feed flows caused by the proportionating valves on the suction side of the pump. This is due to the fact that, during the onset of the suction stroke, said dead volume has first to expand to the volume at ambient pressure before “fresh” liquid can flow into the displacement area. This causes a reduction in the filling stroke efficiency a problem not addressed by the pumps of the prior art discussed above.
Based on the prior art described above, it is the object of the present invention to provide a piston pump as well as a method for controlling said pump, wherein, even with varying specific compressibility of the medium to be delivered and changing delivery pressures of up to 100 MPa and possibly more, the full pumping efficiency is maintained and a (residual) pulsation of the delivery flow is reduced or even suppressed compared to known pumps and methods.
As regards the device, this object is achieved by a piston pump comprising at least two piston/cylinder units for delivering the medium out of a low-pressure area into a high-pressure area, a drive unit for driving at least one piston/cylinder unit, a control unit for controlling the rotational speed of the drive unit and a sensor for collecting an actual value of a control parameter, by which actual value the extent of pulsation of the flow generated on the high-pressure side is derivable, the piston pump being characterized in that the sensor is adapted for collecting values of mechanical forces and/or torques exerted and/or transmitted in/at the structure of at least one piston/cylinder unit and/or of the drive unit.
By this design of the pump according to the invention, the (otherwise) compulsory use of pressure or flow measuring sensors is advantageously avoided. Thus, no sensor has any longer to be arranged in the area wetted by the pumped medium, i.e. in the high-pressure area and the zone wetted by that medium. Rather, it can be arranged in almost any location and is only required to be suitable to detect the mechanical forces and moments exerted and/or transmitted within the structure of the pump during operation of the pump. Preferably, the sensor is arranged such that it is not wetted by the medium to be delivered, especially not by the already pre-compressed or compressed medium.
By structure of the pump basically all mechanical assemblies and components of the pump are meant. In particular, the sensor monitors forces and moments of the aforementioned kind, which are transmitted by or exerted in/at the structure of the piston/cylinder unit, the drive unit or the components interfacing the drive unit and the piston/cylinder unit. The drive unit of the pump in the above sense comprises transmission and drive units. Advantageously, it is a cam drive and comprises associated drive shafts including cam disks as well as the above transmission and drive units.
The described arrangement of the sensor yields several advantages.
By means of this, in comparison with conventionally structured pumps, the transfer volume is significantly reduced, which is decisive for putting low-pressure gradients through the pump, preferably without re-mixing and, consequently, for the suitability thereof in providing a high sample throughput (HTP) by the pertinent analysis system. The transfer volume is defined as the volume in the total liquid duct path ranging from the closing edge of the proportionating valves forming the gradient to the fitting at the outlet of the pump. The smaller the transfer volume, the less time is required for resetting the analysis system to initial conditions for the next analysis run.
If no sensor has to be arranged in the zone wetted by the pumped medium, there is no need for integrating sensors, which are usually two-dimensional rather than desirably tubular-shaped and which also have to be chemically inert depending on their use, which integration in a fluid-tight manner at high pressure is technically complicated and costly. This has the favourable side effect that also the geometries of passages can be configured in a simple manner such that only direct flow-through liquid areas are created.
Moreover, the concept of monitoring by means of force measuring within the mechanical zone not wetted by the pumped medium inherently offers the advantage of a stroke synchronous monitoring of the liquid displacement action, which can be accomplished (even) resolved within a pumping cycle in a functionally robust manner. In contrast to conventional methods which are based on measuring the delivery pressure and the flow rate, delivery pressure artefacts described earlier, which might arise due to a change of the viscosity of the pumped medium or a change of temperature thereof, pressure surges when a sample is fed into the following analysis system, drift of the pressure measuring signal during low pressure gradient operation and clogging of the following column, do not constitute interference factors having a direct effect since not the absolute value of a change but the rate of a change is processed:
Since, with said type of sensor, the measurement of mechanical forces and/or moments exerted/transmitted in/at the mechanical assemblies and components of the piston pump takes the place of a measurement providing only an indirectly measured value (e.g. a direct measurement of the delivery pressure), the accordingly monitored values, apart from controlling the pump, may also be evaluated for further functions, inter alia for checking the (ball) valves at the liquid displacement unit(s) for proper functioning and/or for checking the piston seals for the extent of wear present.
The sensor is preferably one which is adapted to monitor tensile, compression, shear and/or torsional stresses in and/or at the pump structure or components embedded or integrated into it including the drive unit. In principle, this may be a sensor monitoring mechanical forces and/or moments, which especially comprises a strain gauge, a piezo-electric element, an acoustic resonator or an optical measuring device.
For example, a torque or torsion sensor may be used, which is engaged in or at the shaft of the drive unit or cam drive and monitors the torque/torsional moment exerted or transmitted there. Another way is to monitor the torque transmitted at the drive unit of the cam drive by means of the sensor. It is of particular advantage that the sensor is not arranged in the area of the pump wetted by the pumped medium on the high-pressure side. Advantageously, even a wetting of the sensor by the medium on the low-pressure side is also avoided.
The piston pump according to the present invention is preferably a parallel or serial dual-piston pump. In accordance with a particularly advantageous embodiment as a serial dual-piston pump, the piston/cylinder units are formed as a working-piston unit, especially having inlet and outlet valves, and as a storage-piston unit, which are connected to each other in such a way that the working-piston unit draws in the medium to be delivered from the low-pressure area of the pump and provides said medium to one portion of the high-pressure area supplied by the pump on the outlet side and to the other portion on the inlet side of the storage-piston unit in order to achieve a constant delivery flow by means of stroke movements adjusted to each other and such that the storage-piston unit provides the full delivery rate during the subsequent suction stroke of the working-piston unit.
The piston pump preferably comprises one drive unit each for each individual piston/cylinder unit. Moreover, each drive unit can be driven by a separate motor. This results in a high degree of technical freedom with regard to controlling and monitoring the delivery volumes of the individual piston/cylinder units, since both drives may be adjusted to each other within wide limits by modulating the rotational speed of both the motors.
As regards the method, the object of the present invention is achieved by a method for controlling a piston pump for delivering a pumped medium out of a low-pressure area into a high pressure area, preferably a dual-piston pump, which piston pump comprises at least two piston/cylinder units, at least one drive unit for driving at least one of the piston/cylinder units and a sensor for monitoring mechanical forces and/or torques transmitted and/or exerted in/at the structure of at least one piston/cylinder unit and/or the drive unit, which method comprises the following steps:
i) driving at least one piston/cylinder unit by a drive unit at a first speed (n), thereby monitoring mechanical forces and/or torques exerted/transmitted in/at the structure of said piston/cylinder unit(s) and/or said drive unit by the sensor,
ii) monitoring the moment of the actual onset of delivery of the pumped medium out of the piston/cylinder unit(s) into the high-pressure area by using the values of the mechanical forces and/or torques monitored by said sensor,
iii) monitoring the rate of compression at the moment of the actual onset of delivery, particularly the presence of over- or under-compression (so called compressibility compensation),
v) modulating or adjusting the drive unit rotational speed to a second speed (n+1), in such a way that a varying compression rate due to varying system back pressure and/or varying compressibility of the pumped medium arising on the high-pressure side is compensated and in such a way that an essentially pulse-free delivery flow is generated in the high-pressure area,
v) repetition of method steps i) to iv) for each pumping cycle by using the drive unit rotational speed (n+1) modulated in step iv) as the first speed (n).
The method according to the present invention is preferably recursively pursued for each pumping cycle. However, a control cycle can also be based on the measured data of several delivery strokes. The method enables the generation of a high-pressure delivery flow over a wide delivery pressure and specific liquid compressibility range, said high-pressure delivery flow being essentially free of pulsation. In principle, one or more piston/cylinder units can be monitored by sensors. This is to be considered when reference is made solely to one piston/cylinder unit for the sake of convenience throughout the present specification.
When starting-up the pump upon the onset of the pumping operation, i.e. shortly after power on, the pump first has to pass through a start-up phase until a steady delivery flow is reached after its initial start-up. Basically, the drive speed may have an arbitrary value during the start-up phase; it is, however, preferably adjusted to the specific compressibility of the range of pumped media to be expected as well as to the desired rate of delivery, which facilitates the subsequent control of the pump.
After reaching a steady delivery flow or a stable delivery behavior of the pump, the actual method according to the present invention commences. According to this method, the pump is driven at a first speed (n) at least during one pumping or delivery cycle. Preferably, the mechanical forces and/or moments transmitted/exerted in/at the structure of the pump are continuously monitored by the sensor. In particular, forces and/or moments are monitored, which are transmitted and exerted, respectively, in/at the structure of the piston/cylinder units, the drive unit or by intermediate units. In accordance with a particular embodiment, control programs can be employed for this purpose, which are adapted to the respective rate of delivery chosen.
In the subsequent method step ii), the time and thus the piston position is determined by using the mechanical forces and/or moments monitored by the sensor, at which time the actual delivery of the medium into the high-pressure area sets on while driven at the first speed (n) or at which piston position this happens at least at the piston/cylinder unit(s) monitored by the sensor. The onset of the actual delivery can be monitored unambiguously and particularly well by using the course of the mechanical forces and/or moments monitored by the sensor, for example by mathematically deriving the gained measured values or series of measured values, e.g. with respect to time or place position (piston stroke position).
While driving the piston/cylinder unit at the first speed (n), system pressure builds up within the pertinent cylinder displacement chamber as the delivery stroke of the piston increasingly advances. This pressure increases until it is equal to the system pressure prevailing in the adjacent high-pressure area. At the time at which there is a pressure balance within the high-pressure area and within the cylinder space of the piston/cylinder unit, the medium is not yet delivered from the cylinder space. Only after the balance between the pressure within the cylinder space of the piston/cylinder unit and the system pressure on the high-pressure side has been exceeded can the actual delivery take place.
Considering the mechanical forces and/or moments monitored by the sensor, the onset of the actual delivery becomes recognizable in that the course of the monitored force and/or moment measuring values changes significantly. The utilization of the courses of the force and/or moment measuring values monitored by the sensor can be advantageously facilitated by additionally or alternatively assessing and monitoring the first and/or second stroke position dependant derivations of these values or courses.
Since the kinematic characteristics of the drive unit, in particular of drive cam(s) and the (respective) speed (n) of the cam drive(s) are known, the rate of (pre-) compression of medium previously effected within the piston/cylinder unit(s) can be calculated by using the stroke position detected by the sensor or the specific drive unit position (cam angle), at which the actual delivery has set on. In particular, the presence of an over or under compression (over- or under compressibility compensation) can be determined. In other words, it can be determined, whether the (actual) delivery starts too early or too late with regard to the respective geometry of the drive unit (cam drive) present and whether a (residual) pulsation of the generated high-pressure flow arises or may arise as a result. In the case of an angular range integrated in the drive unit or cam drive specially for generating a pre-compression stroke it can thus be determined, whether, at the respective delivery pressure and/or the respective specific compressibility of the pumped medium, the compression caused by employing said cam section at given drive speed over or under compensates the influence of the specific compressibility of the pumped medium on the pumping efficiency, and thus results in a residual pulsation of the delivery flow. This feature is of particular significance when operating the pump according to the low pressure side gradient method (supplying the medium on the suction side with a composition changing as a whole).
Determination of the (pre-) compression effected prior to the onset of delivery is preferably performed in a stroke-related manner. In particular, it is checked whether an over or under compensation of the specific compressibility of the pumped medium is present by using the monitored pulsation characteristics.
In the further process of the method, the drive unit rotational speed (or cam drive speed) is modulated or adjusted to a second speed (n+1). Said speed may be higher or lower than the previous speed (n), which is dependent on the actual required pre-compression determined in the previous pumping cycle. As a result of this modulation, a previously determined too high or too low rate of pre-compression (over or under compensation) is compensated for or its extent is at least minimized. As a consequence, despite of changed system pressure on the high-pressure side and/or a changed specific compressibility of the pumped medium, a delivery flow substantially free of pulsation is generated.
The second speed (n+1), which is adapted to the changed operating parameters, is used as the first speed (n) in step i) in the further process of the method and the method steps described above are recursively pursued. Preferably, this is done already during the respective subsequent piston stroke.
Based on a reference value derived in the described manner, with the described method, it is possible to compensate the influence of changing compression ratios from one stroke to the subsequent stroke, which ratios arise as a result of the changed specific compressibility of the pumped medium and/or the changed delivery pressure caused by the variable system pressure on the high-pressure side.
The drive unit rotational speed (cam drive speed) preferably remains constant during an initial delivery stroke, more preferably during a (initial) delivery stroke initiating the described control method. In particular, the monitoring of the mechanical forces and/or moments or the course thereof is performed during a delivery stroke at a constant drive speed.
As regards the method, it is of particular advantage in the method according to the present invention, if the medium to be delivered is pre-compressed in each pumping cycle. The pre-compression stroke required for that is preferably adapted to the generation of a maximum delivery pressure and an expected maximum specific compressibility of the medium to be delivered. The pre-compression generated by said stroke is sufficient for at least one specified maximum delivery pressure and for an application-specific expected maximum specific compressibility of the fluid to be delivered. Consequently, it is to be expected that during operation at the first speed (n) an over-compression will arise in method step i), whereupon the drive unit rotational speed is reduced by modulation during the pre-compression stroke. This is described as an iterative subtractive approach to pump control elsewhere in this description.
According to a particular embodiment of the method according to the present invention it is foreseen to derive a correction factor by means of the stroke position of the delivery piston determined (by the sensor), in which position the actual delivery sets on. By means of this correction factor a running program adapted thereto is generated or chosen from a plurality of stored running programs, which then forms the basis for the modulation of the drive unit rotational speed, especially within the cam section relevant for the pre-compression stroke.
By means of the described pre-compression and/or modulation of the drive speed (primary control) which results in a modulation of the piston speed, an excessive delivery in relation to the previously chosen nominal delivery amount will take place. In particular, an excessive delivery in relation to a (supplied) volume at ambient pressure will take place as initially; the pump is driven to compensate for maximum compressibility and maximum operating pressure. According to a further suggestion of the present invention, this excessive delivery is compensated for by superimposing a secondary correction factor onto the modulated (cam) drive speed, which correction factor provides for a compensation of the excessive delivery (secondary control: generating true mass flow).
In accordance with a particular embodiment of the method according to the present invention, the values or value patterns detected by the sensors are used to correct the efficiency loss in relation to the suction stroke, which arises at the beginning of each suction stroke by expansion of the detrimental dead volume in the displacement area of the piston/cylinder unit compressed during the previous delivery stroke in accordance with the respective specific compressibility of the pumped medium and in accordance with the respective delivery pressure, i.e. the volume, which, due to the design, is inevitably not displaced during the delivery stroke, however, must be compressed prior to the onset of the actual delivery. This is preferably done by a determination of the duration of the decompression phase. The portion of the stroke volume which cannot be displaced from the displacement area of the piston/cylinder unit during each pumping cycle due to expansion of the detrimental dead volume is monitored. The non-usable portion of the suction stroke results from the expansion of the detrimental dead volume. In particular, during low pressure gradient operation, i.e. when the composition of the medium to be delivered changes and subsequently varying compressibility and/or viscosity affects the back-pressure. This is of particular advantage at very high delivery pressures, since the gradient composition of the medium flowing into the pump during the subsequent suction strokes can be adjusted in accordance with the extent of expansion of the detrimental dead volume. Preferably, this is done by opening and closing the solenoid valves usually employed for this purpose in an adaptively controlled manner at the inlet of the pump for each subsequent suction stroke.

BRIEF DESCRIPTION OF THE DRAWINGS

Further advantages and features of the present invention are apparent from the following description of a non-restrictive embodiment with reference to the figures, in which:

FIG. 1 shows a schematic illustration of the pump according to the present invention including a control unit in an apparatus for (high-pressure and ultrahigh-pressure) liquid chromatography,

FIG. 2 shows diagrams, in which the normalized strokes of the working piston (upper diagram) and of the storage piston (lower diagram), are illustrated as a function of the angle of rotation of the cam shaft,

FIG. 3 shows diagrams, in which the normalized delivery rate of the working piston (upper diagram) and of the storage piston (lower diagram) are illustrated as a function of the angle of rotation of the cam shaft,

FIG. 4 shows diagrams, in which the normalized forces present at the working piston (upper diagram) and at the storage piston (lower diagram), are illustrated as a function of the angle of rotation of the cam shaft,

FIG. 5 shows a diagram, in which the alteration of the force present at the working piston is illustrated as a normalized first derivative, as a function of the angle of rotation of the cam shaft, and

FIG. 6 shows diagrams, in which the open condition of the inlet valve (upper diagram) and of the outlet valve (lower diagram) are illustrated as a function of the angle of rotation of the cam shaft.

FIG. 7 a,b show diagrams, in which for the most preferred embodiment the normalized stroke of the working piston (FIG. 7 a) and of the storage piston (FIG. 7 b), are illustrated as a function of the angle of rotation of the cam shaft,

FIG. 8 a,b show diagrams, in which for the most preferred embodiment the normalized delivery rate of the working piston (FIG. 8 a) and of the storage piston (FIG. 8 b), are illustrated as a function of the angle of rotation of the cam shaft,

FIG. 9 shows a diagram, in which for the most preferred embodiment the normalized cam rotational velocity is illustrated as a function of the angle of rotation of the cam shaft,

FIG. 10 a.b show diagrams, in which for the most preferred embodiment the normalized forces present at the working piston (FIG. 10 a) and at the storage piston (FIG. 10 b), are illustrated as a function of the angle of rotation of the cam shaft,

FIG. 11 shows a diagram, in which for the most preferred embodiment the normalized first derivation of the force at the working piston is illustrated as a function of the angle of rotation of the cam shaft,

FIG. 12 shows the principal layout of the dual stage gear system,

FIG. 13 shows the gear system including one of the two Z-shaped drive arms, and

FIG. 14 shows a lengthwise cut in flow direction of the most preferred embodiment of the liquid end.

DETAILED DESCRIPTION OF EMBODIMENTS

An exemplary embodiment of the piston pump according to the present invention is shown in FIG. 1. The pump 1 designed according to the serial delivery principle comprises a working piston/cylinder unit 2 and a storage piston/cylinder unit 3. The working piston/cylinder unit 2 essentially consists of a working piston 4 performing a reciprocating movement within the cylinder space of a working cylinder 5. Similarly, the storage piston/cylinder unit 3 consists of a storage piston 6 performing a reciprocating movement within the cylinder space of a storage cylinder 7. The working piston 4 and the storage piston 6 are each driven by a separate drive unit. As is apparent from FIG. 1, both drive units are identical, therefore, only the drive unit of the working piston 4 is explained in detail in the following. The explanations equally hold true for the storage piston 6.
The drive unit of the working piston 5 comprises a motor 8. In principle, any kind of motor may be used as motor 8, however, a conventional electric motor, such as a direct-current motor, a magnetostrictive or a piezo-electric drive system is preferred. The motor 8, in combination with an absolute encoder unit 9, is coupled to the gear 10 featuring a drive cam. The position of the drive cam within the gear 10 is monitored and associated in an exact and correctly polarized manner in relation to the stroke movement of the working piston 4 by means of the absolute encoder unit 9. This may also be done, for example, by means of index disks or rotary encoders in combination with specialized control software.
The motor 8 is controlled by a servo unit 11, which in turn is connected to a computing unit 12. A control program and preferably, in table form, specific control programs for modulating or adapting the rotational speed of the motor in accordance with various compression correction factors are stored inter alia in the computing unit 12 or are generated computationally.
Between the gear 10 and the pertinent piston (working piston 4 or storage piston 6) one force/moment sensor 13 each is arranged both in or at the drive of the working piston/cylinder unit 2 and in or at the drive of the storage piston/cylinder unit 3. The force/moment sensor 13 monitors forces and/or moments exerted or transmitted between the respective piston and the transmission unit. The gear unit 10 is configured as a cam drive unit. The cams of said unit basically may have arbitrary kinematic profiles, however, an angular range for generating a pre-compression stroke initiating the delivery stroke and adapted to maximum specified operating conditions with regard to delivery pressure and) to medium compressibility is always provided. Said cam section is employed in accordance with the extent of the system pressure in an aliquot manner by maintaining a basic drive speed, which is modulated in the surplus section not used according to definition such that the stroke movements of the working piston and the storage piston generate a composite constant delivery flow in accordance with the set delivery rate.
The computing unit 12 is connected via the servo unit 11 to both the drive of the working piston/cylinder unit 2 and the drive of the storage-piston/cylinder unit 3. As indicated by the dashed lines in FIG. 1, the computing unit 12 is further connected to the absolute encoder unit(s) 9 as well as to the force/moment sensor(s) 13 and monitors and/or processes the measuring signals thereof.
The working-piston/cylinder unit 2, on its inlet side, comprises an inlet valve 15 at its inlet 14 acting as an uncontrolled check valve. A proportionating valve unit 16 is arranged before said inlet valve being fitted with four special solenoid valves for low pressure side gradient formation. Said unit is, in turn, connected to the computing unit 12 for control purposes and is controlled by said computing unit. During operation of the pump, medium to be delivered can be withdrawn from several respective reservoirs 17 a-d by suitably controlling the proportionating valve unit 16. Gradient formation on the suction side is accomplished by conducting medium alternately from the reservoirs or sources of medium 17 a-d via the inlet valve 15 into the working cylinder 5 by a programmed control of the solenoid valves of the proportionating valve unit 16 during the suction stroke of the working piston 4. During the delivery stroke of the working piston 4 following the suction stroke, the medium drawn into the working cylinder 5 is delivered via an outlet 18 of the working-piston/cylinder unit 2, through an outlet valve 19 also configured as an uncontrolled check valve and via the inlet 21 into the storage piston/cylinder unit 3, which is without valves in the case illustrated.
The delivery flow from the working-piston/cylinder unit 2 to the storage-piston/cylinder unit 3 flows into the storage cylinder 7 via the inlet 21. The stroke movements of the working piston 4 and the storage piston 6 are adapted to each other in such a way that the storage piston 6 simultaneously performs its suction stroke during the delivery stroke of the working piston 4. Accordingly, the working piston 4, on the one hand, delivers into the storage-piston/cylinder unit 3 and, on the other hand, through this unit into the high pressure line 20 forming the feed line to the system supplied with the medium. With the present functional principle, no further compression of medium takes place in the storage piston/cylinder unit 3. The medium is always under system pressure therein. Compression as well as pre-compression is performed exclusively in the working piston/cylinder unit 2. The storage-piston/cylinder unit 3 only serves as a storage and delivery reservoir for bridging the interruption of delivery of the working piston/cylinder unit 2 during its suction stroke. The storage-piston/cylinder unit 3 may form a reservoir having a volume which can be adapted to the respective operating conditions. The volume stored in the storage piston/cylinder unit 3 is delivered from the storage piston/cylinder unit 3 into the high-pressure line 20 during the suction stroke of the working-piston/cylinder unit 2.
In functional continuation of the system, a high-pressure sample injection valve 24 is arranged in the high-pressure line 20 or behind that line in front of a separation column 23. At the outlet of the separation column 23 a detector 28 is arranged, by means of which the chemical compounds can be detected which are injected with the sample volume and being eluted by differential retardation from the separation column by the delivery flow according to their differential partition between the separating phases. Medium conducted through the separation column 23 and the detector is received in a waste container 29.
The pump unit depicted in FIG. 1 further shows, e.g. for safety reasons, a pressure sensor 25 monitoring the medium pressure (system pressure) prevailing in the high-pressure line 20, the values of which are transmitted via the signaling line, illustrated by a dashed line, to the computing unit 12 for [back-up] control purposes. It must be noted that the additional pressure sensor 25 is optional and not compulsory but represents an optional control device.
In FIGS. 2 to 6 exemplary operating diagrams of the piston pump illustrated in FIG. 1 are shown. In each of the diagrams, the angle of rotation of the cam shaft is plotted on the abscissa within a range from 0° to 360°, wherein specific rotation angle positions to be considered are highlighted.
Point A marks the angle of rotation of the cam shaft, at which the pre-compression phase of the working piston 4 is completed. B characterizes the angle of rotation, at which delivery of medium into the high-pressure area is affected solely by the working piston 4. C characterizes the angle of rotation, at which the exclusive delivery of medium solely by the working piston 4 terminates. D characterizes the angle of rotation, at which the medium begins to flow from the low-pressure area into the cylinder space of the working piston/cylinder unit 2 upon onset of the suction phase. Each of the diagrams of FIGS. 1 to 6, on its abscissa, shows a full revolution of the cam shaft (from 0° to)360°. After employing the cam section from the angle of rotation 0 to point E, which corresponds to a complete revolution of the cam shaft over 360°, the entire working cycle commences again at position 0. At which point the delivery stroke of the working piston 4 commences (upper diagram, FIG. 2).
The stroke of the working piston 4 performed from 0 to A generates pre-compression only. Both the inlet valve 15 and the outlet valve 19 remain closed during employing this stroke range (FIG. 6). The storage piston 6 performs part of its delivery stroke within the same rotating range of its drive cam (lower diagram, FIG. 2). As can be seen from FIG. 3, delivery of medium into the high-pressure line 20 over this rotation angle range is solely performed by the storage-piston/cylinder unit 3. The delivery rate achieved by the working piston/cylinder unit 2 is completely consumed for pre-compression (upper diagram, FIG. 3)—the piston displacement action does not deliver pumped medium, it compresses it. FIGS. 4 to 5 show the normalized forces present at the working piston 4 as well as at the storage piston 6 and the normalized alteration of the forces present at the working piston 4 (the first derivative at cam stroke position of the normalized force). The upper diagram of FIG. 4 shows that the force present at the working piston 4 rises in an even and, if necessary, constant manner within the rotation angle range from 0 to A. This may also be gathered from FIG. 5 illustrating an alteration of the force present at the working piston 4.
When point A of the angle of rotation of the cam shaft is reached, the medium within the cylinder space of the working-piston/cylinder unit 2 is pre-compressed in accordance with the system pressure on the high-pressure side. This means that the pre-compression phase is completed. The upper diagram of FIG. 6 shows that the outlet valve 19 opens at this point of time. The stroke of the working piston 4 advances with unchanged speed just as in the pre-compression phase between 0 and A (upper diagram, FIG. 2), whereas the stroke of the storage piston 6 is slightly delayed as compared to the course of the pre-compression phase between 0 and A (lower diagram, FIG. 2). FIG. 3 shows that in the range between A and B the normalized displacement action of the storage piston 6 slightly drops in comparison to that during the pre-compensation phase between 0 and A by the amount of the normalized displacement action of the working piston 4, which corresponds to the representation in the upper diagram. The displacement action of the working piston 4 is no longer consumed for pre-compression in the range between A and B. This results in the fact that both the storage piston 6 and the working piston 4 deliver medium into the high-pressure line 20. As shown in the force diagrams of FIGS. 4 and 5, the normalized force present at the working piston 4 as well as the normalized force present at the storage piston 6 remains constant.
In rotation angle position C, the storage piston 6 has reached the dead center of its stroke movement, the end of the filling stroke. At this point, the changeover from suction stroke to delivery stroke takes place as a result of the reversal of movement (lower diagram, FIG. 2). In the range between B and C, the delivery of medium into the high-pressure line 20 is based on the displacement action of the working piston 4 only.
Between B and C, the stroke of the working piston 4 in relation to the angle of rotation rises considerably compared to the stroke during the pre-compression phase (O-A) and compared to the phase of common delivery of the working piston 4 and the storage piston 6 (A-B) (upper diagram, FIG. 2). As can be seen from the lower diagram of FIG. 2, the suction stroke of the storage piston 6 commences and terminates within that range at B and C, respectively. The delivery forces present at the working piston 4 in the range between B and C correspond to the normalized forces in the range between A and B, since the working piston 4 continues to deliver at system pressure (nominal pressure) on the high-pressure side. The lower diagram of FIG. 4 further shows that, due to the continued exposure to system pressure, the normalized force transmitted at the storage piston 6 remains constant also during the suction stroke. The reason for that being that the storage piston 6 is under system pressure both on the inlet side and on the outlet side. Referred to a full pumping cycle, the force value curve shows the theoretical constant course. It runs parallel to the x-axis since the storage piston is continuously exposed to the full hydrostatic pressure governed by the system pressure prevailing beyond of the pumps outlet.
The additional force induced by the friction between the piston and its seal varies with the system pressure which governs the force pressing the sealing lip of the seal against the piston surface in the contact area. Said friction force, however, reaches only a fraction of the force value which is hydrostatically exerted onto the pump structure being monitored there. The seal friction force is superimposing the hydrostatic force in additive as well as in subtractive manner. Looking to actual curves of the force values monitored during a pumping cycle reveals that the superimposition of the friction force manifests in a (very small) stepwise offset whereby the addition or subtraction depends on the direction of the piston stoke movement: An additive effect—shown in FIG. 4—is associated with the displacement stroke (section O-B and C-E) and a subtractive effect is associated with the filling stroke (section B-C).
It can be seen from FIG. 3 that the normalized displacement action of the storage piston 6 in the range between B and C (during the suction stroke) becomes negative. The normalized displacement action of the working piston 4 exceeds the combined displacement action of both pistons by this amount, said displacement action being illustrated by a dotted line in the diagrams of FIG. 3. In the range between B and C the displacement action of the working piston 4 is divided into the amount of medium volume pumped by the storage-piston/cylinder unit 3 into the high-pressure line 20 and into the amount of medium volume received in the storage-piston/cylinder unit 3 during the suction stroke thereof.
In rotation angle position C, the working piston 4 is at its top dead center-end of displacement stroke—and the storage piston 6 is at its bottom dead center—end of the filing stroke. In this rotation angle position C, the stroke movements of the working piston 4 as well as of the storage piston 6 proceed in the respective opposite direction. This is apparent, in the upper diagram of FIG. 4, from a steep drop of the normalized force present at the working piston 4. The delivery phase of the working piston 4 terminates upon reaching the angle of rotation C, whereas the displacement action of the storage piston 6 commences, wherein, in the latter phase, the displacement action of the storage piston in total corresponds to the combined displacement action of the working piston 4 and the storage piston 6 (FIG. 3). At the angle of rotation C, the outlet valve 19—associated with the working piston—closes and thus prevents a backflow of medium from the high-pressure side into the cylinder space 5 of the working-piston/cylinder unit 2.
The inlet valve 15 opens with delay. This is apparent especially from a comparison of the diagrams of FIG. 6, from which can be gathered that the inlet valve 15 opens only at the angle of rotation D. In the rotation angle range between C and D, the detrimental dead volume, which was not displaced from the working-piston/cylinder unit 2 but being still compressed to system pressure during the previous delivery stroke, expands again to the volume at ambient pressure. Consequently, medium cannot be drawn in from the low-pressure side into the cylinder space 5 of the working-piston/cylinder unit 2 during said expansion phase. Accordingly, the normalized force present at the working piston 4 substantially reaches a value of 0 only in the rotation angle position D.
In rotation angle position D, the negative pressure required for the actual initiation of the suction stroke is established in the cylinder space 5 of the working-piston/cylinder unit 2. In the process the inlet valve 15 opens (upper diagram, FIG. 6) and the suction stroke of the working piston 4 is performed starting at point D until the top dead center is reached at the angle of rotation E. In the range between D and E as well as in the range between C and D the delivery of the medium from the pump into the high-pressure line 20 is based solely on the displacement action of the storage piston 6. In rotation angle position E, which corresponds to the angle of rotation 0, the cycle described above is repeated.
A pressure sensor 25 is arranged in the high-pressure line 20 behind the storage-piston/cylinder unit 3. This sensor is not required for performing the method according to the present invention and can thus, in principle, also be omitted. In the illustrated embodiment, the pressure sensor 25—as shown by the dotted line—is connected to the computing unit 12, preferably via a signalling line. The measured pressure values provided by the pressure sensor 25, for example, permit conclusions as to the cause of failures and malfunctions, including those of the total system 1, by comparing them to measured values provided by the force/moment sensor 13.
In the most preferred embodiment, the invention is carried out on the basis of a dual piston serial type pump 1 comprising a ‘working piston’ 4 and ‘a storage piston’ 6, with said pump 1 having a single DC-motor 8 paired with a belt drive 40 rotating a dual cam axle 41. The rotational motion of the cam disk 42 profiles is converted into linear reciprocating motion by rollers 43 at the near end of Z-shaped drive arms 44. By means of axial ball bearings, the drive arms 44 perform their reciprocating motions along precision guide axles and at the same time are stabilised against canting by means of laterally arranged roller bearings. Furthermore, the drive pistons 4,6, at their distant ends, are each fitted with a hard contact rod 45 in a bracketed holder element 46, with the low friction counter surface of said rod providing, in conjunction with a fork-type spring, a free-floating, no side load inducing connection between the drive and the pumping piston 4,6.
For allowing continuous monitoring of the liquid displacement process in dependency of the delivery pressure and the specific compressibility of the medium being pumped, the working drive arm 44 is fitted with a strain gauge. By means of the strain gauge 49 the efficiency of the pumping and filling stroke can be continuously and precisely monitored during each pumping cycle and, processing its signal track forms the basis for a fast and efficient feedback control method to compensate for the impact of the specific compressibility of the pumped medium at high and highest operation pressures on the pumping performance and efficiency, Compare: FIG. 10 a, b.
The pertinent liquid displacement assembly (LDA) is built to sandwich design according to U.S. Pat. No. 5,653,876.
The kinematic data of the drive cam disks 42 for the working 4 and the storage piston 6 are shown in the figures as diagrams of elevation normalised relative to the maximum stroke amplitude of the working piston 4, covering a full pumping cycle) (360°, FIGS. 7 a and 7 b for working 4 and storage pistons 6 respectively. The pertinent cam reference angles A to E are shown for the implemented cam profiles. These figures should be compared with the idealised plots of FIG. 2.
Covering a full pumping cycle)(360°, the actual normalised delivery rates with the given cam profiles, at constant angular velocity versus cam angular position are shown in FIGS. 8 a and 8 b for working and storage pistons respectively. The pertinent cam reference angles A to E are shown for the actual (implemented) cam profiles. These should be compared with the idealised plots of FIG. 3.
Exemplarily, the actually measured angular velocity as a function of the cam angular position is shown in FIG. 9 for the case of a sufficient compressibility compensation process step during each pumping cycle. Between A and B, i.e. after the appropriate pre-compression level has been reached in A, during the control period, the motor speed is reduced such that the cam rotational speed is lowered to approximately 40% and, subsequently, the cam rotational speed is increased again to its initial nominal value in such way that a composite constant delivery flow is generated by both pistons.
In an ideal way, deceleration is achieved inherently almost instantaneously, since the outlet check valve is hydraulically actuated with negligible delay, whereas the acceleration to the nominal cam velocity up to point B is synchronized with the cam drive profiles. Feedback trigger events in section A to B are processed in the control unit as the first derivative of the monitored force at the working piston, as a function of the angle of rotation of the cam shaft.
The actual normalised forces, measured with the force sensors, are shown in FIG. 10 once a complete compressibility compensation control cycle has been achieved by applying the method of control already described in successive cycles. Typically, 2 to 3 pumping stroke cycles have to be performed until steady and continuous flow conditions are established at the high pressure outlet side of the system.
The shown signal traces are equivalent to the idealised traces in FIG. 4. The characteristic negative force of small value from rotation angle section D to E, measured at the working piston 4, quantifies the friction force of the piston seal free from hydraulic load, with the inlet check valve 38 opened for the filling stroke. Depending on the sign of the piston movement, the storage piston force sensor measures an alternating effective friction due to the contribution caused by the pertinent seal which is permanently under load depending of the system back pressure. The small deviations of the actual measured values relative to the theoretical values are contributed mainly by non-linear variations of the seal friction.
As described previously, the seal friction force is superimposing the hydrostatic force in an additive as well as in a subtractive mode. Looking to actual curves of the force values monitored during a pumping cycle reveals that the superimposition of the friction force manifests by a comparably small stepwise offset whereby addition or subtraction depends on the direction of the piston stoke movement: An additive effect—shown in FIG. 10 b—is associated with the displacement stroke (section O-B and C-E) and a subtractive effect is associated with the filling stroke (section B-C).
The actual normalised first derivative of the measured force is shown in FIG. 11. It is equivalent to FIG. 5. The shown raw and un-dampened signal trace being noisy, still reveals amplitudes which are well above required levels to discriminate the various phases in the pumping cycle.
In a preferred embodiment a Maxon Motor™ RE 268214 with graphite brushes is used. However any DC motor with adequate power rating, sufficient torque and high bandwidth rotational speed specifications matching or exceeding the performance values of that motor will be sufficient to drive the pump system described.
The motor 1 is typically operated within the range of a few rpm up to 12.000 rpm to generate the pump's flow rate range specified from 0 to 5 ml/min maximum.
In FIG. 12 the principal layout of the dual stage gear system is depicted. Using a commercially available gear, it is built according to known design. The dynamic velocity and acceleration range required for generating the specified flow rate range is met by a dual stage belt drive system providing a reduction rate of 100:1. By its concept, the gear system corresponds to a conventional and commercially available design.
FIG. 13 shows the gear system including one of the two Z-shaped drive arms 44 or primary pistons in cross-sectional view. The drive pistons 44 are fitted with axial ball bearings 47 to perform their reciprocating motions along a precision guide axle. They are actuated by the cam profiles 42 via a roller 43 each at their near ends, converting the rotating cam motion into reciprocating piston stroke motions. At their distant end they bear in a clamping end piece 48 a contact rod 45 from hard material which exhibits a low friction counter surface for the pumping piston 44. In conjunction with a fork-type spring, said surface ensures a free-floating connection between the actuation end and the pumping piston shaft 45.
The strain gauge 49 to be seen in the left ‘arm’ of the drive piston 44 is fixed with adhesive into a countersunk hole. In scale, the strain gauge 49 shown is shown oversized for clarity. In the present embodiment, the gear system is identical for both the working 4 and the storage piston 6; hence only one half section of the gear system is shown. Among the possible force sensing devices, the type of force sensor used in the most preferred embodiment is based on the known resistive strain gauge principle. The strain gauges 49 are fixed with cured adhesive and they are conveniently located in the cantilevered part of the Z-shaped drive piston 44, where the flexure and/or shear strains reach their measurable maximum.
A commercially available and industrial standard resistive strain gauge type 49 is used, such as the Vishay™ Type 062LV. Said type allows enables an optimum of shear strain measurement including fully balanced Wheatstone bridge arrangements. The force sensor(s) required for implementing the feedback pump control concept according to the invention is integrated into the structure of the drive piston 44 associated with the working piston 4 and, optionally also in the drive piston 44 associated with the storage piston 6.
Shown in FIG. 14 is a lengthwise cut in flow direction of the most preferred embodiment of the liquid end.

Claims

1. A piston pump for generating a substantially pulsation-free delivery flow of a medium to be delivered, comprising:

at least first and second piston-cylinder-units for delivering the medium out of a low-pressure area into a high-pressure area;

a drive unit for driving the piston-cylinder units;

a control unit for controlling the driven speed of the drive unit; and

a sensor for collecting an actual value of a control parameter, by which actual value the extent of pulsation of the flow generated in the high-pressure area is derivable;

wherein the sensor is adapted for collecting values of mechanical forces and/or torques exerted or transmitted in or at the structure of at least one piston-cylinder unit and/or the drive unit.

2. The piston pump of claim 1, further comprising:

an area wetted by the medium being pumped; and

a dry area, which is not wetted by the medium during operation;

wherein the sensor is located in the dry area of the pump and is not wetted by the medium to be delivered.

3. The piston pump of claim 1, wherein the mechanical forces include at least one of tensile stress, compression stress, shear stress and torsional stress.

4. The piston pump of claim 1, wherein the sensor is based on one of a strain gauge, a piezo-electric element, an optical measuring device and an acoustic resonator.

5. The piston pump of claim 1, wherein the piston pump is a parallel dual piston pump.

6. The piston pump of claim 1, wherein the piston pump is a serial dual piston pump.

7. The piston pump of claim 1, wherein each piston-cylinder unit is driven by a separate drive unit.

8. The piston pump of claim 1, wherein each drive unit is driven by a separate motor.

9. The piston pump of claim 1, wherein the first piston-cylinder unit has a suction side in fluid communication with the low-pressure area and a discharge side in fluid communication with the second piston-cylinder unit and the high-pressure area in parallel.

10. The piston pump of claim 9, wherein a discharge side of the second piston-cylinder unit is hydraulically coupled to the high-pressure area.

11. The piston pump of claim 9, wherein the sensor monitors mechanical forces and/or torques being exerted or transferred in or at the structure or drive of the first piston-cylinder unit.

12. The piston pump according to claim 11, further comprising an additional sensor configured to monitor mechanical forces and/or torques being exerted or transmitted in or at the structure or drive of the second piston-cylinder unit.

13. The piston pump of claim 1, wherein the drive unit is designed to generate a pre-compression stroke.

14. The piston pump of claim 1, wherein the drive unit comprises a cam drive.

15. A method for controlling a piston pump for delivering a medium out of a low-pressure-area into a high-pressure-area, the piston pump including at least first and second piston-cylinder units, a drive unit for driving at least one of the piston-cylinder-units, and a sensor for monitoring mechanical forces and/or torques exerted or transmitted in or at the structure of at least one of the piston-cylinder units and/or said drive unit, comprising:

i) driving at least one piston-cylinder unit by a drive unit at a first speed (n), and monitoring using the sensor mechanical forces and/or torques exerted or transmitted in or at the structure of the at least one driven piston-cylinder unit and/or the drive unit;

ii) monitoring the moment of the actual onset of delivery of the pumped medium out of the at least one driven piston-cylinder unit into the high-pressure-area by using the values of the mechanical forces and/or torques monitored by the sensor;

iii) monitoring the rate of compression of the pumped medium at the moment of the actual onset of delivery;

iv) modulating the drive unit rotational speed to a second speed (n+1), such that a varying compression rate due to varying system back pressure and/or varying compressibility of the pumped medium arising on the high-pressure side of the at least one driven piston-cylinder unit is compensated and a substantially pulse-free delivery flow of the pumped medium is generated in the high-pressure area and

v) repeating steps i) to iv) for each pumping cycle using the drive unit rotational speed (n+1) modulated in step iv) as the first speed (n).

16. The method of claim 15, wherein in step ii) the variation of the values measured for the mechanical forces and/or torques is monitored.

17. The method of claim 15, wherein the mechanical forces and/or torques or the course of their measuring values in step ii) are monitored during a single pump stroke at constant drive unit rotational speed.

18. The method of claim 15, wherein stored running programs are used for modulation of the drive unit rotational speed.

19. The method of claim 18, wherein a correction factor is derived based on the actual onset of delivery and wherein a specific rotational speed control program is selected depending on the correction factor.

20. The method of claim 15, wherein in step i) a rotational speed control program is applied, which is tuned to the flow rate set at the piston pump.

21. The method of claim 15, wherein the medium being pumped is pre-compressed at the beginning of each pumping cycle.

22. The method of claim 21, wherein a secondary correction factor is superimposed onto the drive unit rotational speed which compensates for the excess delivery flow referred to volume at ambient pressure generated by the precompression.

23. The method of claim 21, wherein the pre-compression stroke is adapted to a specified maximum delivery pressure and a maximum specific compressibility to be expected for the pumped medium.

24. The method of claim 21, wherein the drive unit rotational speed is reduced when being modulated during the pre-compression stroke.

25. The method of claim 15, wherein the composition of medium to be delivered by the pump varies over time.

26. The method of claim 15 wherein a volume which cannot be expelled from the displacement area of the piston-cylinder unit of the working piston as detrimental dead volume, but still has to be compressed according to current operation conditions before actual pumping can onset, is monitored for each pumping cycle in order to determine the loss of filling efficiency due to expansion of said volume at the beginning of the suction stroke.

27. The method of claim 25, wherein the gradient composition for the subsequent filling stroke is corrected according to the expansion of the not dischargeable detrimental dead volume within the displacement area of the first piston-cylinder unit.

28. The method of claim 15, wherein the drive unit comprises a cam drive.