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EP1841965A1 - Compresseur a piston axial - Google Patents

Compresseur a piston axial

Info

Publication number
EP1841965A1
EP1841965A1 EP06706406A EP06706406A EP1841965A1 EP 1841965 A1 EP1841965 A1 EP 1841965A1 EP 06706406 A EP06706406 A EP 06706406A EP 06706406 A EP06706406 A EP 06706406A EP 1841965 A1 EP1841965 A1 EP 1841965A1
Authority
EP
European Patent Office
Prior art keywords
compressor
suction
pressure
compressor according
control device
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Withdrawn
Application number
EP06706406A
Other languages
German (de)
English (en)
Inventor
Otfried Schwarzkopf
Ullrich Hesse
Henrick Brandes
Oliver Tschismar
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Valeo Compressor Europe GmbH
Original Assignee
Valeo Compressor Europe GmbH
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Valeo Compressor Europe GmbH filed Critical Valeo Compressor Europe GmbH
Publication of EP1841965A1 publication Critical patent/EP1841965A1/fr
Withdrawn legal-status Critical Current

Links

Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B27/00—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders
    • F04B27/08—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders having cylinders coaxial with, or parallel or inclined to, main shaft axis
    • F04B27/14—Control
    • F04B27/16—Control of pumps with stationary cylinders
    • F04B27/18—Control of pumps with stationary cylinders by varying the relative positions of a swash plate and a cylinder block
    • F04B27/1804—Controlled by crankcase pressure
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B49/00—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00
    • F04B49/22—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00 by means of valves
    • F04B49/225—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00 by means of valves with throttling valves or valves varying the pump inlet opening or the outlet opening
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B27/00—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders
    • F04B27/08—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders having cylinders coaxial with, or parallel or inclined to, main shaft axis
    • F04B27/14—Control
    • F04B27/16—Control of pumps with stationary cylinders
    • F04B27/18—Control of pumps with stationary cylinders by varying the relative positions of a swash plate and a cylinder block
    • F04B27/1804—Controlled by crankcase pressure
    • F04B2027/1863—Controlled by crankcase pressure with an auxiliary valve, controlled by
    • F04B2027/1872—Discharge pressure
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B27/00—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders
    • F04B27/08—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders having cylinders coaxial with, or parallel or inclined to, main shaft axis
    • F04B27/14—Control
    • F04B27/16—Control of pumps with stationary cylinders
    • F04B27/18—Control of pumps with stationary cylinders by varying the relative positions of a swash plate and a cylinder block
    • F04B27/1804—Controlled by crankcase pressure
    • F04B2027/1863—Controlled by crankcase pressure with an auxiliary valve, controlled by
    • F04B2027/1877—External parameters
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B2205/00—Fluid parameters
    • F04B2205/01—Pressure before the pump inlet
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B2205/00—Fluid parameters
    • F04B2205/17—Opening width of a throttling device
    • F04B2205/171—Opening width of a throttling device before the pump inlet
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B2207/00—External parameters
    • F04B2207/04—Settings
    • F04B2207/044—Settings of the rotational speed of the driving motor

Definitions

  • the invention relates to an axial piston compressor, in particular a compressor for the air conditioning system of a motor vehicle according to the preamble of claim 1.
  • An example of such an axial piston compressor is known from DE 197 49 727 A1.
  • This comprises a housing and arranged in the housing, driven by a drive shaft compressor unit for sucking and compressing a refrigerant.
  • the compressor unit is controlled substantially by the pressure P c in an engine chamber depending on the load or the rotational speed of the compressor, wherein also a suction pressure P V1 and a high pressure P V2 , which prevail on the suction or outlet side of the compressor influence take on the control of the compressor.
  • the regulation takes place via a change in the piston stroke of the compressor, which is determined by the deflection of a swash plate from a zero position.
  • Swing plate heard (including a moment due to the center of gravity, this proportion may be positive, for example, according to DE 195 14 748 C2, that is (+))
  • Pistons, sliding blocks or even an oscillating swash plate can belong.
  • the moment due to the rotating masses is generally effective over a wide range of tilt angles. Only in the range of very small tilt angle can be generated for example by an exposed center of gravity ( Steiner share in the calculation of the moment of deviation J y2 ) at the swash plate a aufregelndes moment. Furthermore, DE 195 14 748 C2 a course for the moment due to translationally moving masses indicated. This is, as already explained, effective.
  • EP 0 809 027 A1 discloses a compressor in which an attempt is made to compensate the delivery rate of the compressor by the dynamic behavior of the compressor engine, so that the delivery rate of the refrigerant or the refrigerant mass flow can be kept constant.
  • the flow rate at varying rotational speeds is excited to take advantage of the restoring torque of the swash plate, which counteracts their inclination due to dynamic forces on the co-rotating disc part.
  • DE 103 29 393 A1 which is assigned to the Applicant, further explains why the component mass should not be the preferred parameter in order to influence the control behavior of the engine as a result of speed fluctuations as desired.
  • the desired control behavior of the compressor is achieved according to DE 103 29 393 Al primarily not with the component mass of the swash plate in relation to the translationally moving masses, but taking into account the mass moment of inertia of the arrangement of the swash plate, which depends more on the geometry of the same their component mass.
  • a key idea within this application is to compensate or overcompensate for the moment due to translationally moving masses in the event of speed fluctuations or speed changes directly as a result of the rotating mass.
  • a compressor 101 of the type described is regulated.
  • Such a compressor provides a suction gas pressure level as well as a high pressure level during operation.
  • the refrigerant circuit has these pressure levels.
  • a certain pressure adjustment or a pressure adjustment takes place by means of a Expansionsotgans 103, which in turn reacts to changes in the operating state of the compressor and optionally intervene regulating.
  • a pressure is set, for example by control valves on the compressor, which is between the Sauggasdruckieri and the high pressure level.
  • the engine room pressure engages in the balance of forces and the moment balance on the swash plate such that the tilt angle of the swash plate can be adjusted.
  • the swashplate is adjusted to maximum tilt angle. If the engine room pressure is set significantly above the suction pressure, the swashplate is adjusted to a minimum tilt angle.
  • the regulation is effected by the possible volume flows between the individual chambers or pressure layers.
  • the reference numerals 102 represent a
  • Gas cooler / condenser, 104 an evaporator and 105 a controlled system.
  • a controlled system 105 dashed lines which has P V2 as the setpoint.
  • Such a controlled system is more common especially for the refrigerant CO 2 .
  • the model described here is simplified and to be regarded as exemplary.
  • a compressor with a compressor unit according to the preamble of claim 1 has an additional control and / or control device for the suction gas side refrigerant mass flow and / or the pressure on the suction side and / or the suction density.
  • an embodiment of the present invention is applied to an axial piston compressor, it being understood that this should by no means be construed restrictively, since a construction according to the invention may also be applied to a whole series of compressors of a different type. If the refrigerant flows into the cylinders of the axial piston compressor at a lower pressure level or a lower suction density, the compressor will have a tendency to be depleted, ie to reduce the piston stroke, at a constant pressure in the engine chamber. Accordingly, the limitation of the refrigerant mass flow or the suction pressure engages directly in the control of the compressor. In other words, a change in the volume flow leads in the main to a change in the pressure difference regulating the compressor and thus to a predetermined control characteristic.
  • the control device comprises a throttle point with an actuator.
  • the throttle point may be a throttle valve or a throttle valve.
  • a pressure reducer is conceivable.
  • the actuator controls the refrigerant mass flow or the suction pressure acting on the piston (P V1 *).
  • the actuator of the control device sets the refrigerant mass flow or the suction pressure as a function of the speed.
  • Speeds are an easily accessible controlled variable. A detection of rotational speeds, for example, by the generation of electrical impulses (induction principle) take place, but it is, for example, a direct, centrifugal force-dependent control conceivable. By a speed-dependent control are so opened various constructive embodiments of a compressor according to the invention, with low production costs offer an advantage here as well.
  • the throttle point preferably has a stop associated with the actuator for a position of minimum flow cross-section, wherein this is arranged such that even at very high rotational speeds of the compressor, a predetermined minimum refrigerant mass flow or suction pressure is ensured. This ensures in a simple manner that the compressor does not automatically completely de-regulate even at very high speeds due to the throttle point or the additional control device.
  • the actuator is in a simple and thus inexpensive to produce variant of a compressor according to the invention, a control piston, which may be present in particular in the form of a stepped piston.
  • control device for controlling or limiting the suction-gas-side refrigerant mass flow or the suction pressure or the suction density.
  • a control device generally has a structurally simple construction.
  • the control device comprises at least one inlet valve arranged on the suction gas side.
  • the control device may be integrated in the inlet valve (s).
  • the inlet valve is a pressure-controlled flap valve, which is formed in a structurally simple variant by a valve plate with through-throttle bore and a suction plate.
  • the suction lamella is preferably tongue-shaped.
  • the compressor according to the invention is a compressor with a piston, in particular an axial piston compressor which has a cylinder block and at least one, but in particular 5 to 9 pistons, which can be axially moved back and forth in bores provided in the cylinder block, then optionally one cylinder can be used Be associated intake valve, wherein the corresponding suction lamellae for the cylinder are integrated in a Sauglamellenplatine. This reduces the required number of items for a compressor according to the invention, which reduces the manufacturing cost.
  • the inlet valve associated with or facing the end of the or each cylinder chamber comprises in a further preferred embodiment, an annular extension which may be bevelled or flattened towards the attachment point of the suction plate.
  • the stroke of the suction plate can be effectively limited.
  • the ratio of piston diameter and piston stroke is about 0.4 to 1.5, in particular about 0.65 to 1.1.
  • the ratio of piston diameter and passage throttle bore in the valve plate is preferably in about 1.5 to 5, in particular 2.5 to 4.
  • the ratio of passage throttle bore in the valve plate and the stroke of the suction lamella is in a further preferred imple mentation form about 2 , 5 to 8, especially 3.7 to 6.7.
  • the ratio of piston stroke to the stroke of the suction plate can be about 10 to 30, in particular 14 to 24.
  • control device may also include a suction-side arranged aperture, which defines the refrigerant mass flow or suction pressure or the suction density accordingly.
  • the compressor according to the invention is a compressor which has a swashplate
  • the deflection angle of the swashplate which determines the piston stroke of the compressor, is largely determined by the interaction of the pressure P c in an engine compartment essentially accommodating the swashplate and the refrigerant mass flow determined on the suction side or the suction pressure P V i * on the other hand.
  • Another force acting on the piston is generated by the pressure P V2 on the high pressure side.
  • the control device can be actuated or controlled from outside the compressor. In particular, this is a magnet coil or the like. Device in question. Simple maintenance and a simple replacement of the control device for the control device is guaranteed.
  • the regulating and / or control device comprises a suction-gas-side oil separator having multifunctional significance. On the one hand, oil which is present in the suction gas is separated, on the other hand, pressure regulation or regulation of the refrigerant mass flow can thereby simultaneously take place.
  • control device is self-regulating and in particular dependent on the difference of the pressures on the outlet or high pressure side on the one hand and the inlet and suction side of the compressor on the other. This ensures a safe control of the compressor, taking into account the most important operating parameters.
  • a torque distribution of the rotational and the translationally movable or movable components of the compressor is so pronounced that an increase in the compressor speed ensures a substantially constant control characteristic (ie balanced moments).
  • Torque ratio causes the compressor to be neutral in relation to the speed, ie in particular not alsregelt.
  • a Abregeltul is supported if necessary by a corresponding control intervention for the pressure P c in the engine room, in particular, smaller control interventions on the adjustment or defined control tion / limitation of the suction pressure P V1 * can be done without significant power losses of the engine.
  • the regulating and / or control device can be arranged in a suction channel extending primarily in the cylinder head of the compressor.
  • the suction duct connects a suction gas inlet of the compressor with a suction chamber upstream of and upstream of the inlet openings of the individual cylinders.
  • the control device may further comprise means for measuring the Kälteschmas senstrom and / or the pressure in the suction duct (both sides of the throttle point and on one side of the throttle point) and / or the speed of the compressor and / or the motor driving and / or a Compressor output side pressure include.
  • the control device may further comprise means for measuring the Kälteschmas senstrom and / or the pressure in the suction duct (both sides of the throttle point and on one side of the throttle point) and / or the speed of the compressor and / or the motor driving and / or a Compressor output side pressure include.
  • the control device may further comprise means for measuring the Kälteschmas
  • the actuator of the control device optionally acts against the force of an elastic element, in particular against the force of a spring.
  • This allows a wide range of different control characteristics (depending on which characteristic has the elastic element) and at the same time represents an easy-to-implement automatic control of the actuator.
  • the force of the elastic element in particular by a set screw or the like Device adjustable. This ensures that with one and the same structure different control characteristics can be easily adjusted. Tolerances in the production of the elastic element or the properties of the elastic element can thus be easily compensated, since a fine adjustment of the characteristic of the elastic element is possible.
  • the actuator is arranged between the compressed gas and the suction gas side and is thus regulated by the application of the pressures on the one hand the compressed gas side and on the other hand, the suction gas side automatically (possibly against the action of the elastic element).
  • the regulating or control device comprises a throttle point with a constant cross section.
  • This throttle point can be present, for example, as the only regulating device in a compressor according to the invention or else in combination with a throttle point, which comprises an actuator.
  • the desired result can already be achieved by means of this simple structural measure.
  • the suction line and / or a connection between the compressor and an evaporator are part of the control device, wherein in particular in the case of a throttle point with a constant cross-section efficient control of the compressor by a corresponding design of the suction line and / or the connection between the compressor and the Evaporator can be achieved.
  • a particularly efficient variant of a compressor according to the invention is obtained when a pressure difference of about 1 bar at a compressor speed of about 600 U / min and / or of about 10 bar at about 8000 U / min is present on both sides of the control device.
  • the control device or throttle point preferably comprises a pipe having a pipe cross-section of about 8 mm to 10 mm, and this ensures a desired control characteristic in particular in the case of a throttle point with a constant cross-section.
  • CO 2 is used as the refrigerant.
  • a particularly efficient and with few control interventions concerning the pressure P c in the engine room ausumbleder compressor results when a moment due to the rotary moving components of the compressor is substantially equal to a moment due to the translationally moving components of the compressor ie when the compressor in behaves speed-neutral in its control behavior.
  • Fig. 1 shows a first preferred imple mentation form of an inventive
  • Fig. 2 is a schematic representation of the operation of a second preferred imple mentation of a compressor according to the invention
  • Figs. 10 to 13 are further examples of control characteristics of a compressor having a tendency to be up-regulated as the number of revolutions increases, and a compressor exhibiting a decelerating tendency;
  • Fig. 14 is a schematic detail view of a third preferred embodiment of a compressor according to the invention.
  • a first preferred embodiment of a compressor comprises a housing 1, a cylinder block 2 and a cylinder head 3.
  • pistons 4 are mounted axially movable back and forth.
  • the drive of the compressor via a pulley 5 by means of a drive shaft 6.
  • a compressor with variable piston stroke wherein the piston stroke is determined by a deflection angle of a swash plate 7.
  • the swash plate 7 is via sliding blocks 8 with the piston 4 in operative engagement and is rotationally driven by the drive shaft 6.
  • the deflection angle of the swash plate 7 can, as known from the prior art, by a pressure change in an engine chamber 8 ', in which substantially the swash plate is arranged, to be influenced.
  • the engine chamber 8 ' may be pressurized between a suction pressure, that is, a pressure prevailing at an inlet side of the compressor, and a high pressure, i. a pressure prevailing at an outlet side of the compressor, be applied.
  • a suction pressure that is, a pressure prevailing at an inlet side of the compressor
  • a high pressure i. a pressure prevailing at an outlet side of the compressor
  • the deflection angle of the swash plate 7 influencing variable is the torque distribution between the translationally movable components of the cylinder, such as the piston 4 or the sliding blocks 8, and the rotationally movable components of the compressor, such as the swash plate 7.
  • the compressor at high speeds has a rather abregelde tendency. This is especially desirable in modern compressors to avoid icing, especially at high speeds without a large number of control actions.
  • a suction gas channel 9 is arranged in the cylinder head 3, which connects a suction gas inlet 10 with a suction gas chamber 11, which is arranged upstream of the cylinders.
  • the compressed fluid or refrigerant is provided to the refrigerant circuit via a compressed gas or outlet chamber 12.
  • a control device is mounted to regulate the suction gas-side refrigerant mass flow and thus also the pressure on the suction side of the compressor.
  • This comprises a control piston 13, which may be formed as a stepped piston, alternatively to the illustrated embodiment, an elastic element in the form of a spring 14 and a set screw 15, which serves to adjust the bias of the spring 14.
  • the adjusting piston 13 is acted upon on its side of the outlet or pressure gas chamber 12 inclined side with the outlet or high pressure, while on the adjusting screw 15 inclined side, i.
  • the suction gas inlet 10 inclined side is acted upon by the suction or inlet pressure.
  • the actuator piston 13 is associated with a stop for a position minimum flow cross-section. This ensures that even at very high speeds of the compressor and a relatively high output pressure, a predetermined minimum refrigerant flow or suction pressure is ensured.
  • the control unit with the actuating piston 13, the spring 14 and the screw 15 is thus self-regulating, the control is done depending on the pressures at the outlet and the inlet or suction gas side.
  • the deflection angle of the swash plate 7 is determined by the interaction of the pressure in the engine compartment 8 'on the one hand and the refrigerant mass flow on the suction gas side or the suction pressure on the other hand, the suction pressure itself is in turn dependent on the output pressure of the compressor, so that a feedback control for the compressor is created.
  • the control device which in the schematic representation of FIG. 2 generally serves as a throttle point 17 is referred to, of course, by external control variables as well as by external devices, such as a magnetic coil, are regulated.
  • Fig. 2 it is shown that the throttle body 17 and the throttle (actuator) is controlled by an external signal 16.
  • This signal can be based on a measurement of the mass flow, the pressure on the high pressure side or a differential pressure between the suction gas duct and the high pressure side or a differential pressure in the suction gas channel, which results from the different pressures P V1 and P V1 * on the two sides of the throttle point 17 , to be generated.
  • Other parameters such as a speed, or even temperatures or the like. Sizes as the basis of the signal 16 are conceivable.
  • Fig. 2 is also a schematic representation of the refrigeration cycle in a h vs. , log p diagram (supercritical process, with CO 2 as refrigerant) shown in a representation at the throttle point ( ⁇ PV).
  • the torque distribution between the translationally moving masses of the compressor, such as the piston 4, and the rotationally moving masses, which include, for example, the swash plate 7, has a regulating effect on the compressor.
  • something closer to this moment ratio something closer to this moment ratio.
  • the moments of inertia and moments of deviation, as well as other variables influenced by the geometry and density of the materials, can easily be calculated by means of CAD.
  • J y2 -j a cos ⁇ sin ⁇ + J 3 cos ⁇ sin ⁇
  • M kjges m k R ⁇ Z j cosß
  • the (tilt) moment of the swashplate can be deliberately adjusted as a result of the associated deviation moment by various parameters (geometry, density distribution, mass, center of mass) such that
  • Fig. 4 shows the derivation of the underlying example of the engine of the swash plate design.
  • the tilting moment M sw is simplified as a result of the deviation moment J y2 of the swashplate relative to the translationally moved masses or the moment M K> ges produced thereby.
  • forces and moments of the pins or the gas force support or the like are not detected in the calculation scheme. These are of secondary importance.
  • FIGS. 5, 6 and 7 each show a calculation scheme according to the equations used. Furthermore, the moment equilibrium is displayed as the calculation result. For this purpose, a (tilting) characteristic is also presented (qualitatively), as it would result considering the gas forces.
  • the tilting characteristics of Figures 5, 6 and 7 arise when, in addition to the variation of speed and the engine room pressure in addition to the forces and moments described a certain suction pressure and a certain high pressure system are impressed. In this case, it is assumed that the suction pressure applied in front of the compressor and the high pressure applied to the compressor correspond approximately to the suction pressure and the high pressure in the compressor, ie, there is no throttling in the compressor.
  • FIG. 5 shows an engine with an adjusting behavior
  • FIG. 6 shows an engine with regulating behavior
  • M k ges is approximately equal to M sw , which leads to a speed-neutral control behavior of the compressor. If desired, M sw greater than M k ges can be selected, which leads to a behavior of the compressor at high speeds; however, it is by no means desirable that M k ges be greater than M S ⁇ V (increasing the speed of the compressor).
  • M k ges is approximately equal to M S ⁇ T .
  • the tilt angle of the swash plate is essentially only by varying the pressures P V1 (suction pressure), P V2 (high and outlet pressure), and the engine room pressure p c changed.
  • P V1 suction pressure
  • P V2 high and outlet pressure
  • the change normally occurs essentially only by the engine room pressure p c .
  • a throttle point is provided which is variable and engages quickly.
  • the compressor in such a way that the throttling engages directly as a function of the compressor speed (as in the second preferred embodiment, cf. FIG. 2).
  • the cross-section of the throttle point is a function of the compressor high-pressure P V2 , ie the throttling is controlled as a function of high pressure.
  • the pressure P V1 which prevails before the throttle point, approximately maintains its level (systems ei term s pressure level on the suction side), while the pressure P V i * after the throttle point decreases compared to P V1 .
  • the pressure P V2 now acts as a significant control variable (in addition to the suction pressure) on the throttling mechanism, that the cross-section of the throttle point is reduced.
  • thermodynamic quantities in front of the compressor and behind the compressor remain the same and the control element in the system does not intervene.
  • thermostatic expansion organ In addition to a thermostatic expansion organ, of course, differently operating and differently controlled expansion organs are conceivable.
  • the compressor regulates automatically, in addition to the engine room pressure P c, the intake pressure P V1 or the high pressure P V2 is regulating effect. Since P V1 and P V2 are also paused by the operating state of the system, where as a rule changes are not necessarily also desired when the compressor speed changes, a pressure P V1 * is brought about which acts as a gas force on the piston in the force or moment equilibrium of the swash plate can intervene.
  • the pressure P V1 * can be significantly smaller than P Y1 (by 5 to 15 bar). Since depending on the operating point such a throttling can be associated with significant losses, the throttle point or control device is variable in a wide range.
  • the throttle point which more or less narrows the Sauggas effetsquerrough depending on the position, in a preferred embodiment, three different workspaces:
  • the adjusting piston In a third position, the adjusting piston can fall into a stop when there is a minimal flow cross-section in the suction line. A minimum flow cross-section is maintained (operating position 3).
  • FIG. 9a shows the adjustable mass flows (qualitative representation) for a given pressure level P V1 and P V2 of the air-conditioning system, while FIG. showing pV diagram.
  • the envelope of the corresponding slope shows the delivered mass flow for maximum tilting tilt angle / maximum geometric displacement.
  • the desired mass flow of refrigerant such as m l5 m 2 or m 3, so upon further increase in speed does not increase the refrigerant mass flow is desired.
  • the horizontal curves for m 15 m 2 and m 3 shown in the diagram each correspond to a certain engine room pressure , which is approximately constant. In the area of horizontal lines, the effect of the intake-gas-side throttle point comes into play as speed increases. While on the envelope the throttle-side throttle is in position 1 (no throttling) in operation s, in operation range 2 the throttle angle is reduced with increasing speed.
  • the mass flow can be kept constant.
  • the operating state for the pressures on the high-pressure side and on the suction side of the system and for the intake state at the compressor inlet with P V1 , t vi and P V2 is fixed and is in the operating state
  • the speed n 2 at a refrigerant mass flow in, before, so is the throttle point in the operating state 2 ie the intake cross-section of the suction line is reduced in the region of the throttle point relative to the initial state (operating state 1).
  • a suction pressure P V i * has been set, which is smaller than the pressure P V1 as a result of the throttling.
  • the control behavior is particularly characterized in that, contrary to the prior art where, in a constant operating state of the system described by P V1, vi t, Py 2 each engine chamber pressure P c is exactly one Schwenkusionnkippwinkel is associated (see. Fig. 5, wherein exceptions Range very high speeds or very small tilt angle (maxima) occur, for a engine room pressure P c multiple Schwenkusionnkippwinkel are conceivable.
  • P c is a manipulated variable, but also the pressure P V1 *.
  • the pressure difference P V1 * -P c can reach negative values.
  • the prior art must be based on P C -P V1 .
  • the pressure P c here is always greater than the pressure P V1 .
  • the control range is also larger ( ⁇ p).
  • an external actuation of an actuating piston or a throttle device can also be effected (by a magnetic coil or the like, see FIG.
  • Such a device must be "communicated” as a signal, the mass flow increase, for example by detecting a suction-side or high-pressure side pressure difference (orifice / orifice (variable or non-variable)) on the suction or high pressure side of the compressor).
  • the third preferred embodiment it is a compressor which has no control device but a control device for the suction pressure. This results in that the compressor described is structurally very simple and therefore cost-effective to manufacture. It should be noted at this point that the control device of the third preferred embodiment may be implemented together with a regulator for the suction pressure in a compressor. Alternatively, however, a construction is conceivable which has only one control device.
  • the third preferred embodiment comprises, like the other preferred embodiments, a plurality of pistons 4, which are mounted in the cylinder block 2 back and forth.
  • the third preferred embodiment at the inlet side for the suction gas in the cylinder chamber on a valve plate 18 with a suction plate attached thereto 21 on.
  • the suction plate 21 is tongue-shaped and serves to control the Sauggaseinlasses.
  • the suction lamella 21 closes a through-flow bore 19, while the suction lamella 21 moves downwards during suction of the suction gas (due to the negative pressure prevailing in the cylinder) by a stroke t (indicated by arrows 20) and to be sucked in Refrigerant or the suction gas through the passage throttle bore 19 granted inlet to the cylinder.
  • the passage throttle bore 19 has a diameter d. Due to the geometry of the intake valve, that is to say in particular due to the diameter d of the passage throttle bore 19 and the compressor geometry, over a wide operating range of the compressor according to the invention a desired lowering of the pressure P V1 * occurs. This can be achieved, for example (in the case of a compressor with refrigerant CO 2 ) with the following parameters for the compressor geometry:
  • the stroke t of the suction plate 21 is between 0.9 and 1.2 mm, while the valve plate 18 has a bore (through-throttle bore), the diameter d is between 4.5 and 6 mm.
  • the values for the piston diameter are approximately 15 to 18 mm and the piston stroke is approximately 17 to 22 mm.
  • the maximum displacement per cylinder is 3 ccm to 6 cc.
  • the energetically favorable variables describing the geometry of the compressor are a ratio of piston diameter and piston stroke of about 0.65 to 1.1, a ratio of piston diameter and passage throttle bore in the valve plate of about 2.5 to 4, a ratio of passage throttle bore in the valve plate and stroke of the suction lamella of about 3.7 to 6.7 and a It should be noted at this point that these values reflect the optimum geometry for operation with CO 2 as a refrigerant, but that, depending on design requirements, also values from 0.4 to 1 , 5 for the piston diameter to piston stroke ratio and values from 1.5 to 5 for the piston diameter and throat ratio, and 2.5 to 8 for the ratio of throat and stroke of the suction fin and values of about 10 to 30 for the ratio of piston stroke to the lift of the suction plate are energetically favorable.
  • the through-flow bore 19 on the suction side is therefore used on the suction side as a throttle point or as a control device and designed specifically in conjunction with the other parameters regulating the compressor.
  • a construction is very effective in compressors, which are already torque optimized, ie, which have an optimal ratio between the moments due to the rotational masses and due to the translationally moving masses.
  • the inflowing gas flows through a suction chamber, which is mounted in the cylinder head 2, with the pressure P V1 and is then introduced via the inlet valve, which has, for example, the configuration described above, in the cylinder bore, where due to the Saugventil configuration of the pressure P VJ * which ensures optimum control behavior of the compressor.
  • the throttling of the suction pressure or the refrigerant flow generates a Abregel bin, which results not primarily from a reduction in the suction density, but from the direct use of the resulting pressure difference across the throttle for Hubvolumenver too.
  • By adjusting the throttle results in an adjustment of the resulting pressure at the throttle pressure difference and thus an adjustment of the stroke volume.
  • a change in the volume flow leads to a change in the resulting pressure difference and thus to a readjustment of the stroke volume.

Landscapes

  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Compressors, Vaccum Pumps And Other Relevant Systems (AREA)
  • Control Of Positive-Displacement Pumps (AREA)

Abstract

L'invention concerne un compresseur, notamment un compresseur pour l'installation de climatisation d'un véhicule automobile, qui comprend un carter (1) ainsi qu'une unité de compression placée dans ce carter (1), entraînée par un arbre d'entraînement et servant à aspirer et à comprimer un fluide réfrigérant, l'unité de compression étant régulée par la pression (P<SUB>c</SUB>) régnant dans une chambre de mécanisme d'entraînement délimitée pratiquement par le carter (1). Le compresseur selon l'invention est caractérisé en ce qu'il comprend un dispositif de régulation (17) et/ou de commande supplémentaire pour le flux massique de réfrigérant côté gaz aspiré, pour la pression d'aspiration et/ou pour la densité d'aspiration.
EP06706406A 2005-01-25 2006-01-25 Compresseur a piston axial Withdrawn EP1841965A1 (fr)

Applications Claiming Priority (3)

Application Number Priority Date Filing Date Title
DE102005003494 2005-01-25
DE102005007849A DE102005007849A1 (de) 2005-01-25 2005-02-21 Axialkolbenverdichter
PCT/EP2006/000646 WO2006079525A1 (fr) 2005-01-25 2006-01-25 Compresseur a piston axial

Publications (1)

Publication Number Publication Date
EP1841965A1 true EP1841965A1 (fr) 2007-10-10

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ID=36072227

Family Applications (1)

Application Number Title Priority Date Filing Date
EP06706406A Withdrawn EP1841965A1 (fr) 2005-01-25 2006-01-25 Compresseur a piston axial

Country Status (5)

Country Link
US (1) US20080138212A1 (fr)
EP (1) EP1841965A1 (fr)
JP (1) JP2008536035A (fr)
DE (1) DE102005007849A1 (fr)
WO (1) WO2006079525A1 (fr)

Families Citing this family (7)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO2007068335A1 (fr) * 2005-12-17 2007-06-21 Ixetic Mac Gmbh Compresseur de climatiseur
DE102006056823A1 (de) * 2006-12-01 2008-06-05 Valeo Compressor Europe Gmbh Verdichter
DE102008008355A1 (de) 2008-02-08 2009-08-13 Valeo Compressor Europe Gmbh Verdichter
DE102009038462A1 (de) * 2009-08-21 2011-03-03 Dürr Systems GmbH Taumelkolbenpumpe zur Dosierung eines Beschichtungsmittels
DE202011110410U1 (de) * 2011-06-01 2014-03-18 Continental Reifen Deutschland Gmbh Kompressor mit Druckbegrenzung
BR102017010629B1 (pt) * 2017-05-19 2024-04-30 Nidec Global Appliance Brasil Ltda Compressor hermético de deslocamento positivo
DE102021105945B4 (de) 2021-03-11 2022-12-01 Bayerische Motoren Werke Aktiengesellschaft Verfahren und Erkennungseinrichtung zum Erkennen einer Vereisungssituation einer Klimaanlage sowie Kraftfahrzeug

Family Cites Families (27)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPS51105608A (fr) * 1975-03-13 1976-09-18 Sentoraru Jidosha Kogyo Kk
JPH0637874B2 (ja) * 1984-12-28 1994-05-18 株式会社豊田自動織機製作所 可変容量圧縮機
JPS62253970A (ja) * 1986-04-25 1987-11-05 Toyota Autom Loom Works Ltd 可変容量圧縮機
JPS6365177A (ja) * 1986-09-05 1988-03-23 Hitachi Ltd 可変容量斜板式圧縮機
JPH0343684A (ja) * 1989-06-08 1991-02-25 Toyota Autom Loom Works Ltd 可変容量圧縮機
JP3417652B2 (ja) * 1994-04-21 2003-06-16 株式会社豊田自動織機 容量可変型斜板式圧縮機
JP3505233B2 (ja) * 1994-09-06 2004-03-08 サンデン株式会社 圧縮機
JP3561366B2 (ja) * 1996-03-29 2004-09-02 サンデン株式会社 強制リデュース装置及びそれを備えた圧縮機
JPH10325393A (ja) * 1997-05-26 1998-12-08 Zexel Corp 可変容量型斜板式クラッチレスコンプレッサ
JP2000009034A (ja) * 1998-06-25 2000-01-11 Toyota Autom Loom Works Ltd 空調システム
JP4181274B2 (ja) * 1998-08-24 2008-11-12 サンデン株式会社 圧縮機
JP2000145629A (ja) * 1998-11-11 2000-05-26 Tgk Co Ltd 容量可変圧縮機
JP2000146312A (ja) * 1998-11-13 2000-05-26 Toyota Autom Loom Works Ltd 空調装置
JP2000230481A (ja) * 1999-02-10 2000-08-22 Toyota Autom Loom Works Ltd 容量可変型圧縮機のクランク圧制御機構
JP2001132632A (ja) * 1999-11-10 2001-05-18 Toyota Autom Loom Works Ltd 容量可変型圧縮機の制御弁
JP2001221157A (ja) * 2000-02-04 2001-08-17 Toyota Autom Loom Works Ltd 可変容量圧縮機
DE60122225T2 (de) * 2000-02-18 2007-07-12 Calsonic Kansei Corp. Taumelscheibenkompressor mit variabler Verdrängung
JP3933369B2 (ja) * 2000-04-04 2007-06-20 サンデン株式会社 ピストン式可変容量圧縮機
US6439199B2 (en) * 2000-04-20 2002-08-27 Bosch Rexroth Corporation Pilot operated throttling valve for constant flow pump
JP3837278B2 (ja) * 2000-08-10 2006-10-25 株式会社神戸製鋼所 圧縮機の運転方法
JP2002349440A (ja) * 2001-05-30 2002-12-04 Zexel Valeo Climate Control Corp 往復式冷媒圧縮機
EP1394412B1 (fr) * 2001-06-06 2007-03-07 TGK Co., Ltd. Compresseur a deplacement variable
JP3985507B2 (ja) * 2001-11-22 2007-10-03 株式会社豊田自動織機 斜板型圧縮機
JP4023351B2 (ja) * 2002-05-29 2007-12-19 株式会社デンソー 揺動斜板型可変容量式圧縮機
JP2004340007A (ja) * 2003-05-14 2004-12-02 Toyota Industries Corp 可変容量型圧縮機におけるバイパス装置
EP1493923A3 (fr) * 2003-07-03 2006-11-15 Kabushiki Kaisha Toyota Jidoshokki Compresseur à plateau en biais
JP4479504B2 (ja) * 2004-04-28 2010-06-09 株式会社豊田自動織機 可変容量圧縮機

Non-Patent Citations (1)

* Cited by examiner, † Cited by third party
Title
See references of WO2006079525A1 *

Also Published As

Publication number Publication date
DE102005007849A1 (de) 2006-08-17
JP2008536035A (ja) 2008-09-04
US20080138212A1 (en) 2008-06-12
WO2006079525A1 (fr) 2006-08-03

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