Patent Abstract:
The invention relates to a reciprocating piston pump having a magnet drive and a first displacement chamber and a second displacement chamber, which are separated from each other by a piston, wherein both displacement chambers are connected to each other by a fluid-conducting channel. An overflow valve is arranged in the channel and allows a preferred flow from the first displacement chamber to the second displacement chamber, wherein an additional return valve is arranged either in a transition region between an inlet and the first displacement chamber or in a transition region between the second displacement chamber and an outlet, an armature of the magnet drive being firmly connected to the piston.

Full Description:
REFERENCE TO RELATED APPLICATIONS 
     This application is a continuation of International application number PCT/EP2012/000837 filed on Feb. 27, 2012, which claims priority to German application number 10 2011 012 322.9 filed on Feb. 25, 2011. 
    
    
     FIELD 
     The invention relates to a reciprocating-piston pump which is driven by magnets, and to a method for producing and for operating a reciprocating-piston pump. 
     BACKGROUND 
     Reciprocating-piston pumps which are driven by a magnet are known, for example, from documents DE 43 28 621 C2, DE 102 27 659 B4, DE 10 2006 019 584 B4 or DE 10 2008 010 073 B4. Said pumps are used as a rule as metering or delivery pumps and serve to deliver a proportional conveying flow depending on the frequency of the electric actuation. 
     Furthermore, units which are called a metering pump or linearly driven pumps are known, for example, from property rights DE 40 35 835 A1, DE 10 2008 013 441 B4 or DE 298 21 022 U1. 
     DE 35 04 789 A1 describes a reciprocating-piston pump having an electromagnetic drive, in which an armature with a piston, which is connected thereto and configured as a piston rod, is moved away from an outlet on account of the excitation of a coil. The pump also includes a restoring spring which is supported against the armature and a spring abutment being stressed during the movement away from the outlet. When the coil is de-energized, the restoring spring moves the actuator which is formed from the armature and piston rod against an outlet stop which forms an adjustable end stop for the actuator within the housing of the pump. The pump has a suction-side first displacement space which is called a suction space and a second displacement space which is called an armature space, which displacement spaces are connected to one another by a fluid-conducting channel and a nonreturn valve provided therein and radial holes in such a way that a preferred flow from the first to the second displacement space is made possible. Here, a further nonreturn valve is arranged in a transition region between an inlet and the first displacement space. Here, the restoring spring has a prestress which is sufficient to displace the actuator against the outlet upon de-energization and to eject the entire volume of the second displacement space. In addition, the active force of the restoring spring is further reinforced by virtue of the fact that the inlet-side end face of the piston which faces the first displacement space is loaded with fluid there and is therefore pressed in the direction of the outlet. Although the prestress of the restoring spring can also be increased by way of the setting of the position of the outlet stop, its force is already far higher than a counterforce which results from the setpoint value of the pressure and cross section of the outlet face, with the result that no adaptation to the setpoint value of the pressure in the outlet is possible in this way. 
     SUMMARY 
     In one embodiment a pump is provided that generates, not a predefined delivery flow, but rather a predefined pressure at the pump outlet and adapts the delivery flow automatically depending on the requirement of the connected consumer. Since the inlet pressure is known and is approximately constant, the generation of a predefined pressure difference between the outlet and inlet is also expedient. 
     Automatically pressure-regulating pumps are known as rotationally operating pumps from the specialist field of oil hydraulics, to be precise either as valve-controlled variable displacement pumps, for example “Bosch Rexroth A10VOxDR/5”, or as variable displacement pumps, the effective displacement volumes of which are modified directly by the pressure to be regulated, for example “Bosch Rexroth PV7-2X/ . . . ”. The rotary pumps are widespread, but considerably too large and too expensive for the application here. 
     Pressure regulation is also achieved by the combination of a known metering pump with a pressure limiting valve which is connected to the line between the pump and the consumer, but this leads to a higher structural outlay, the risk of oscillations and possibly a considerable temperature influence on the pressure regulation. 
     In one embodiment a reciprocating-piston pump is disclosed having a magnetic drive and a method for producing and operating it, which achieve favorable and reliable automatic pressure regulation with a low structural outlay. 
     According to one embodiment, a reciprocating-piston pump which is driven by a magnet and has the indicated means is designed in such a way that it delivers only the fluid flow which is necessary to maintain the required pressure. To this end, the generated pressure counteracts the movement of the delivery piston and, if the limit value which is predetermined by the force balance at the piston is exceeded, brings the movement of the piston to a standstill. As a result, the piston covers only a part stroke; and the magnitude of the part stroke is dependent directly on the pressure which is built up and indirectly on the fluid requirement of the consumer. 
     In order to utilize the equilibrium of the forces at the piston to regulate the pressure, it is not appropriate, however, to utilize the force of the magnet during the delivery phase, because the magnetic force is subject to great fluctuations as a result of the supply voltage and the coil temperature. Instead, the force of the restoring spring is utilized for delivery and for force calibration. The piston stroke after the magnet is switched on is used merely to pump fluid from the first displacement space into the second displacement space and to stress the restoring spring. The force of the restoring spring is not influenced by the stated disturbance variables of supply voltage and temperature, but rather is dependent substantially on the spring prestress of the restoring spring and the piston stroke. The influence of the stroke can be kept small by the selection of a low spring stiffness, and the pressure to be regulated by the pump can be set by the modification of the spring prestress. 
     If the prestress of the restoring spring can be adjusted only with unacceptable outlay or with risks for the function, it may be suitable to allow a further spring to act on the piston, the prestress of which further spring can be set considerably more easily. It is immaterial here whether the further spring, the so-called correction spring, acts in the same direction on the piston as the restoring spring, or counteracts the restoring spring, as long as only the effects of both springs are dependent on the stroke of the piston and, in the case of opposed action, the force of the restoring spring is greater than the force of the correction spring. 
     The restoring spring or the spring group which comprises the restoring spring and the correction spring produce, as a result of their spring stiffness, a small influence of the stroke on the pressure at the outlet, which influence can be measured, however, and can possibly be utilized. Here, above all, the partial stroke at the end of the delivery phase has an effect on the pressure over averaged time. 
     The described pressure regulation can be realized by way of different known designs of reciprocating-piston pumps, as long as only the delivery of the fluid takes place in the restoring phase of the work cycle, that is to say when the magnet is switched off. The reciprocating-piston pump will as a rule comprise two valves; these can be an inlet valve and an overflow valve between the displacement spaces, or an overflow valve and an outlet valve. 
     In a first embodiment, the reciprocating-piston pump comprises an inlet valve and an overflow valve, and the piston is mounted in the cone in a sliding and dynamically sealing manner. Since the restoring spring is supported in the cone, it is advantageous here not to set the prestress of the restoring spring, but rather to set the prestress of an additional correction spring by means of a displaceable bush. The bush is to be secured after the displacement; this can be achieved by a sufficient interference fit or by welding, soldering, adhesive bonding or calking. 
     In a second embodiment, the reciprocating-piston pump comprises an overflow valve and an outlet valve, and the piston is mounted in the yoke in a sliding and sealing manner. Since the cone does not comprise a sliding bearing for the piston in this case, it is possible here without risk to set the prestress of the restoring spring by means of a displaceable spring bearing. In this case, the stop bush within the spring bearing which represents the inlet-side stop for the piston has to be set subsequently to its correct size, without displacing the spring bearing further. Both the spring bearing and the stop bush have to be secured after the setting operation, in order that they are not displaced further during operation of the pump. A sufficient interference fit, welding, soldering, adhesive bonding or calking can serve to this end. 
     In one embodiment the spring bearing seals the pump to the outside and a completely impermeable seal toward the cone is therefore required; the methods of welding, soldering and adhesive bonding can be used for this purpose, or an elastomer seal can be inserted. 
     For both embodiments, the setting of the restoring spring can also be realized by virtue of the fact that the restoring spring is mounted on one side or both sides on adjusting shims which are selected as required and then inserted after a suitable test operation of the pump or a subassembly. However, this solution is considered to be less advantageous because the described test operation cannot be combined with the final test of the pump after its production. 
     It is also conceivable to set the bush in order to set the spring prestress of the correction spring or the spring bearing not by displacement, but rather to provide the components and the components which enclose them with threads and to perform the setting by way of rotation of the bush or the spring bearing. In this case, the securing of the position will be performed in a known manner by locking with a further component which is provided with a thread or by adhesive bonding. These procedures are also considered to be less advantageous, since they are associated with higher costs and because the seal of a spring bearing which is screwed in is firstly necessary and secondly complicated. 
     In some fields of application of the pump, it is required that, after the pump is switched off, the fluid flows back slowly into the storage reservoir which is connected to the inlet side. To this end, a deliberate leak is then provided in the two valves, which leak is so great that a sufficient outflow takes place after the pump is switched off, but is only so small that the delivery function is not impaired during normal operation. The sealing gap of the dynamic seal between the piston and the piston bearing is also designed for the same leak. 
     In other fields of application, it is required that, after the pump is switched off, a defined residual pressure is maintained, but is not exceeded as a result of temperature-induced expansion of the fluid. To this end, the piston of the pump is provided with an outlet-side sealing stop ring, the active sealing face of which in interaction with the force of the restoring spring results in the required residual pressure. 
     In many applications, an outlet pressure of the pump which is as uniform as possible is required, which outlet pressure is additionally not to be exceeded or is to be exceeded only slightly if the fluid freezes after the pump is switched off. To this end, a compensation volume which is variable under pressure is separated from the second displacement space, which compensation volume is integrated into the pump housing in one advantageous embodiment and therefore requires only a small amount of additional installation space. The variable compensation volume is delimited by a tubular elastic diaphragm; a closed gas volume is situated on that side of the diaphragm which faces away from the working fluid. Fluid dampers are known per se, but not in interaction with pressure-regulating reciprocating-piston pumps as described here. 
     The reciprocating-piston pump according to this invention is distinguished by a very small overall size and low production costs in comparison with known pumps with a similar function. On account of its robustness, it can also be used under adverse environmental conditions in a large temperature range. It is suitable, in particular, for large-scale applications in automotive engineering, for example for the supply of systems for injecting additive or fuel into the exhaust gas section of internal combustion engines. Liquids which freeze in the range of the environmental conditions which are specified for the application can also be conveyed by way of the pump when they have thawed again. 
     Further advantages, developments, properties, features and functions of the invention result from the following description of example embodiments and the claims. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       In the following text, the invention will be explained in greater detail using example embodiments with reference to the appended drawings. 
         FIG. 1  shows a first example embodiment of a reciprocating-piston pump according to the invention in a non-energized state with an inlet valve, without an outlet valve and with a correction spring. 
         FIG. 2  shows a second example embodiment of a reciprocating-piston pump according to the invention without an inlet valve, with an outlet valve, without a correction spring with an adjustable spring bearing for a restoring spring. 
         FIG. 3  shows a third example embodiment of a reciprocating-piston pump according to the invention with a protective means against backflow. 
     
    
    
     DETAILED DESCRIPTION 
       FIG. 1  shows a first example of a reciprocating-piston pump  1  which is driven by a magnet which comprises a magnet housing  2 , a coil  3 , a yoke  4 , a cone  5  and an armature  6 . The primary air gap, at which the axial magnetic force is built up, is situated between the armature  6  and the cone  5 . The secondary air gap between the yoke  4  and the armature  6  builds up only a negligibly small axial magnetic force; the secondary air gap serves only to guide the magnetic flux. 
     The armature  6  is connected to the piston  7  of the pump  1 , and both are pressed into a starting position by a restoring spring  8 . The piston  7  and the armature  6  are additionally loaded with a stroke-dependent force by a correction means which is configured as a correction spring  22 . 
     The magnet is supplied cyclically with the working voltage by an electric actuation means (not shown); the working cycle of the pump  1  is produced by the switching on and off of the working voltage. 
     The piston  7  is mounted in a bore of the cone  5 ; the piston  7  and the cone  5  form a sliding bearing  20  with the cylindrical faces which slide in one another, which sliding bearing  20  is of such tight design that it at the same time also fulfils the function of a dynamic seal with a sealing gap  21 . 
     The interior of the pump  1  is divided into two displacement spaces by the dynamic seal  20 : the first displacement space  25  is connected via an inlet valve  14  to an inlet  13  of the pump  1 ; when the piston  7  is situated in the rest position without magnetic force and pressure, the second displacement space  26  is connected to an outlet  19  of the pump  1 . 
     The two displacement spaces  25 ,  26  are connected to one another by the channel  28  which can run, for example, in the interior of the piston  7  and which comprises an overflow valve  9  which, in one embodiment, permits only a fluid flow from the first displacement space  25  to the second displacement space  26 . 
     In one embodiment the overflow valve  9  is advantageously configured as a ball check valve, comprising a ball  10 , a valve spring  12  and a sealing seat  11  which is part of the piston  7 . Here, the sealing seat  11  is provided with a groove or an elevation which is dimensioned in such a way that a defined leakage flow can flow. 
     The inlet valve  14  is configured as a conical nonreturn valve; it comprises a valve cone  15 , a valve spring  16  and a sealing seat  17  which is part of the cone  5 . 
     In the rest position without magnetic force and pressure, the piston  7  bears via the stop ring  24  against the rear wall of the yoke  4 . In this embodiment, the stop ring is perforated, in order that the channel  28  is always connected to the outlet  19 . 
     The outlet  19  is formed integrally on the yoke  4  and comprises the correction spring  22  which is clamped between a setting bush  23  and the stop ring  24 . 
     The valve cone  15  of the inlet valve comprises a hole (not shown in detail in  FIG. 1 ) which penetrates the valve cone  15  and has a small diameter, as is shown in  FIG. 3  as hole  18 , with the result that a defined leakage which causes a restricted outflow of the fluid toward the inlet  13  is achieved. 
     Finally, the dynamic seal  20  between the piston  7  and the mounting in the cone  5  also has a leak which is dependent on the gap height in the bearing. The gap height is adapted to the leakage requirement in the application. 
       FIG. 1  also describes the integration of a fluid damper into the reciprocating-piston pump  1 . To this end, a diaphragm  27  divides the second displacement space  26 ; that side of the diaphragm  27  which faces away from the fluid is loaded by a gas which is situated in a shut-off space. 
     The function of the pump  1  according to  FIG. 1  can be described best using the temporal sequence: in the rest state which is characterized by a very low pressure at the outlet  19  of the pump  1  and by a de-energized state of the magnet coil  3 , the restoring spring  8  presses the piston  7  onto the outlet-side stop in the yoke  4 . If the magnet coil  3  is then energized, a magnetic force is built up at the primary air gap between the armature  6  and the cone  5 , which magnetic force is greater than the sum of the spring forces of the restoring spring  8  and the correction spring  22 . As a result, the armature  6  and the piston  7  which is connected to it move to the suction side of the pump. The first displacement space  25  is reduced in size, and the pressure therein rises above the pressure of the inlet  13 . As a consequence, the inlet valve  14  closes and the overflow valve  9  opens. Fluid from the first displacement space  25  flows over into the second displacement space  26 . No delivery into the outlet  19  has yet taken place during this stroke. The restoring spring  8  is stressed, and the correction spring  22  is relieved. 
     When the piston  7  reaches the inlet-side stop in the cone  5 , or when the coil current is switched off beforehand, the forward movement of the armature  6  comes to a standstill. As soon as the magnetic force is lower than the sum of the forces of the restoring spring  8  and the correction spring  22 , the movement direction of the armature  6  and of the piston  7  which is configured as a piston rod reverses. The volume of the second displacement space  26  is reduced and the volume of the first displacement space  25  is increased. The pressure in the first displacement space  25  drops and, as a result, the inlet valve  14  opens and fluid flows from the inlet  13  into the first displacement space  25 . 
     The pressure in the second displacement space  26  rises slightly and, as a result, the overflow valve  9  closes. From this instant, fluid is pushed out of the second displacement space  26  into the outlet  19 . 
     Since only a comparatively small fluid quantity is tapped off by the consumer on the outlet side, the pressure in the outlet  19  rises until the pressure limit value which is predefined by the forces of the springs  8  and  22  and the active area of the piston  7  is reached. When the pressure limit value is reached, the movement of the piston  7  comes to a standstill since there is no longer an excess of force in the movement direction. If further fluid is tapped off in this situation by the consumer, the springs  8  and  22  correspondingly continue to press the piston  7  and the pressure changes only slightly in the process. The pump  1  remains in this situation until a new electric actuating signal is issued to the magnet. 
     A new pump cycle begins with the new actuating signal, as described above, but from the position of the piston which was reached last. When the magnet is switched on, the armature  6  and piston  7  move as far as the inlet-side stop and, when the magnet is switched off, they move during operation as intended only as far as the position, in which the spring forces and the pressure force are in equilibrium. This results in a part stroke operation, in which the stroke and therefore the delivery output of the pump are dependent on the requirement of the consumer which is connected downstream and the pressure at the outlet changes only to a small extent which can, however, be influenced by the frequency of the actuating pulses. 
     An alternative example embodiment of a reciprocating-piston pump  101  is shown in  FIG. 2 . The same designations as in  FIG. 1  or the designations incremented by 100 denote the same or structurally comparable parts here which will no longer be introduced separately. 
     In the embodiment according to  FIG. 2 , no inlet valve is arranged in the inlet  13  and, in contrast, an outlet valve  130  is provided in the outlet  19 , which outlet valve  130  ensures the pump function in interaction with the piston  7  and an overflow valve  109 . The outlet valve  130  comprises a ball  31 , a sealing seat  32  and a spring  35 . The outlet valve  130  according to  FIG. 2  has a sealing seat  32  which is provided with a suitable groove or a suitable elevation, in order to make a leakage flow possible. 
     A correction spring  22  is not provided in the embodiment according to  FIG. 2 ; an adjustable spring bearing  29  is provided instead which makes an adjustment of the prestressing force of the restoring spring  8  possible. The adjustable spring bearing  29  and the inlet  13  are configured as one component which can be fixed in the cone  5 . A stop bush  36  which limits the stroke of the armature  6  is situated within the inlet  13 . 
     In contrast to  FIG. 1 , the piston  7  in the embodiment according to  FIG. 2  is mounted in a corresponding bore in the yoke  4 , with the result that the outer circumference of the piston  7  and the bore in the yoke  4  together form a sliding bearing  120  with a sliding seal  121 . 
     Finally, the dynamic seal  120  between the piston  7  and the mounting in the yoke  4  also has a leak which is dependent on the gap height in the bearing  120 . The gap height is adapted to the leakage requirement in the application. 
     A slightly modified function results for the refinement of the pump  101  with an outlet valve  130  and without a correction spring  22  according to  FIG. 2 : in the rest state which is characterized by a very low pressure at the outlet  19  of the pump  101  and by a de-energized state of the magnet coil  3 , the restoring spring  8  presses the piston  7  onto the outlet-side stop in the yoke  4 . If the magnet coil  3  is then energized, a magnetic force is built up at the primary air gap between the armature  6  and the cone  5 , which magnetic force is greater than the force of the restoring spring  8 . As a result, the armature  6  and the piston  7  which is connected to it move to the suction side of the pump  101 . The second displacement space  126  is increased in size, and the pressure therein falls below the pressure of the outlet  19 . As a consequence, the outlet valve  130  closes and the overflow valve  109  opens. Fluid from the first displacement space  125  flows over into the second displacement space  126 . No delivery into the outlet  19  has yet taken place during this stroke. The restoring spring  8  is stressed. 
     When the piston  7  reaches the inlet-side stop on the stop bush  36 , or when the coil current is switched off beforehand, the forward movement of the armature  6  comes to a standstill. As soon as the magnetic force is lower than the force of the restoring spring  8 , the movement direction of the armature  6  reverses. The volume of the second displacement space  126  is reduced and the volume of the first displacement space  125  is increased. The pressure in the first displacement space  125  drops and, as a result, fluid flows from the inlet  13  into the first displacement space  125 . The pressure in the second displacement space  126  rises slightly and, as a result, the overflow valve  109  closes and the outlet valve  130  opens. From this instant, fluid is pushed out of the second displacement space  126  into the outlet  19 . Since only a comparatively small fluid quantity is tapped off by the consumer on the outlet side, the pressure in the outlet  19  rises until the pressure limit value which is predefined by the force of the restoring spring  8  and the active area of the piston  7  is reached. When the pressure limit value is reached, the movement of the piston  7  comes to a standstill because there is no longer an excess of force in the movement direction. If further fluid is tapped off by the consumer in this situation, the spring  8  continues to press the piston  7  correspondingly, and the pressure changes only slightly in the process. The pump remains in this situation until a new electric actuating signal is issued to the magnet. 
     A new pump cycle begins with the new actuating signal, as described above, but from the position of the piston which was reached last. When the magnet is switched on, the armature  6  and piston  7  move as far as the inlet-side stop and, when the magnet is switched off, they move during operation as intended only as far as the position, in which the spring forces and the pressure force are in equilibrium. This results in a part stroke operation, in which the stroke and therefore the delivery output of the pump are dependent on the requirement of the consumer which is connected downstream and the pressure at the outlet changes only to a small extent which can, however, be influenced by the frequency of the actuating pulses. 
       FIG. 3  describes an embodiment of a reciprocating-piston pump  201  which is modified only slightly in comparison with the reciprocating-piston pump  1  from  FIG. 1 , with the result that the same designations as in  FIG. 1  or the designations incremented by 200 denote the same or structurally comparable parts here which will no longer be introduced separately. 
     The reciprocating-piston pump  201  has a stop ring  224  which prevents a further flow of fluid to the outlet  19 , as a result of the sealing of the displacement space  26  with respect to the outlet  19  after the pump  201  is switched off, and maintains a low minimum pressure in the line which is connected at the outlet  19 , which minimum pressure results from the force of the restoring spring  8  and the active sealing area of the stop ring  224 . In this embodiment, the channel  28  is connected to the second displacement space  26  by a hole  233 . 
     A leakage hole  18  which penetrates the valve member  215  axially is shown by dashed lines in the valve member  215  which has the valve cone  15 . 
     A method for pressure setting then takes place as follows: 
     each of the above-described pumps  1 ,  101 ,  201  is assembled in a known way and inserted into a function test bench. The inlet  13  is connected to a supply tank and the outlet  19  is connected to a pressure reservoir. 
     The pump  101  is then energized cyclically and a pressure builds up in the pressure reservoir. The pressure is compared with a setpoint value, and a correction value for setting the spring prestress of the restoring spring  8  is calculated from the deviation of the pressure from the setpoint value. In accordance with the correction value, the spring bearing  29  of the restoring spring  8  is displaced. The spring bearing  29  is gripped with an interference fit in the cone  5  of the magnet, that is to say can be displaced with high force, but then remains in its position during operation of the pump  101 . If the design of the interference fit makes it necessary, the spring bearing  29  is secured after the setting operation. After the setting and securing of the spring bearing  29 , the stop bush  36  is set to its correct size, without displacing the spring bearing  29  further in the process. The bush  36  is also secured if this is required. 
     As an alternative, the pump  1 ,  201  has an additional correction spring  22 , with the result that the spring prestress of the restoring spring  8  does not need to be adjusted. In this case, instead of a spring bearing of the restoring spring  8 , the setting bush  23  is displaced which forms the spring bearing of the correction spring  22 . The setting bush  23  is also gripped in an interference fit, in the component outlet  19  in this case. If it is necessary according to the design, the setting bush  23  is secured after the setting operation. 
     Whereas the above-described setting of the pressure takes place immediately after production, a small change in the pressure can still be achieved during operation, by the frequency of the actuation and therefore the part stroke which is present over the averaged time being changed, because the spring stiffness of the restoring spring and possibly of the correction spring brings about a slightly stroke-dependent force.

Technology Classification (CPC): 5