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EP2021631B1 - Procédé de régulation du débit massique de fluide frigorigène d'un compresseur - Google Patents

Procédé de régulation du débit massique de fluide frigorigène d'un compresseur Download PDF

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Publication number
EP2021631B1
EP2021631B1 EP07725250A EP07725250A EP2021631B1 EP 2021631 B1 EP2021631 B1 EP 2021631B1 EP 07725250 A EP07725250 A EP 07725250A EP 07725250 A EP07725250 A EP 07725250A EP 2021631 B1 EP2021631 B1 EP 2021631B1
Authority
EP
European Patent Office
Prior art keywords
compressor
ges
moment
sub
approximately
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.)
Not-in-force
Application number
EP07725250A
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German (de)
English (en)
Other versions
EP2021631A1 (fr
Inventor
Ulrich Hesse
Oliver Tschismar
Otfried Schwarzkopf
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
Priority claimed from DE102006029875A external-priority patent/DE102006029875A1/de
Priority claimed from DE102006029874A external-priority patent/DE102006029874A1/de
Application filed by Valeo Compressor Europe GmbH filed Critical Valeo Compressor Europe GmbH
Publication of EP2021631A1 publication Critical patent/EP2021631A1/fr
Application granted granted Critical
Publication of EP2021631B1 publication Critical patent/EP2021631B1/fr
Not-in-force legal-status Critical Current
Anticipated expiration legal-status Critical

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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
    • 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
    • F04B27/10—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 having stationary cylinders
    • F04B27/1036—Component parts, details, e.g. sealings, lubrication
    • F04B27/1054—Actuating elements
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B49/00—Arrangement or mounting of control or safety devices
    • F25B49/02—Arrangement or mounting of control or safety devices for compression type machines, plants or systems
    • F25B49/022—Compressor control arrangements
    • 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
    • F04B2201/00—Pump parameters
    • F04B2201/12—Parameters of driving or driven means
    • F04B2201/1201—Rotational speed of the axis
    • 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
    • F04B2205/05—Pressure after the pump outlet
    • 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
    • F04B2205/09—Flow through the pump
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2309/00—Gas cycle refrigeration machines
    • F25B2309/06—Compression machines, plants or systems characterised by the refrigerant being carbon dioxide
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/07—Details of compressors or related parts
    • F25B2400/076—Details of compressors or related parts having multiple cylinders driven by a rotating swash plate

Definitions

  • the present invention relates to a method for controlling a refrigerant mass flow of a compressor according to claim 1 and to a compressor according to the preamble of claim 10.
  • Compressors for automotive air conditioning systems and method for controlling the same are known from the prior art.
  • the refrigerant mass flow of these compressors is generally determined by the lifting height of the pistons of the compressor, wherein the lifting height is defined by the deflection of a pivotable in their relative position to a drive shaft of the compressor inclined or swash plate.
  • the regulation of the deflection angle takes place via a variation of the engine compartment, which is essentially delimited by a housing of the compressor and in which the swashplate mechanism is also mounted.
  • a compressor described therein provides (as already mentioned explains) in operation, a suction gas pressure level p s and a high pressure level p d ready. This is done by (usually) a control valve, through which the operating point is set. Likewise, the air conditioner about these pressure levels.
  • an expansion device which regulates the circulation, pressure separation), which in turn reacts to changes in the operating state of the compressor and optionally intervenes regulating.
  • a pressure p c is set by control valves on the compressor, which is between the Sauggasdruckrise p s and the high pressure level p d .
  • the change of the engine room pressure p c engages in the force or torque balance on the swash plate so that the tilt angle of the swash plate can be adjusted. If the pressure p c in the engine room approaches the suction pressure p s , then the swash plate is adjusted in the direction or to a maximum tilt angle. If an engine room pressure p c is set significantly above the suction pressure p s , then the swash plate is adjusted to a lower or minimum tilt angle.
  • the regulation is effected by the possible volume flows (volume flow 1 between p d and p c , volume flow 2 between p c and p s ) between the individual chambers or pressure layers.
  • a feedback between the compressor (referred to as system) and the control valve, which controls the engine room pressure p c , necessary to achieve a desired refrigerant mass flow can.
  • the control can be due to the mass flow, due to the pressure difference between the high pressure side (p d ) and the suction pressure side (p s ) and also only after the suction pressure (p s ).
  • the control valve (usually a magnetic control valve) is operated via the control variable delivered in the feedback loop and the corresponding desired pressure p c is set in the engine room.
  • a magnetic control valve is operated via the control variable delivered in the feedback loop and the corresponding desired pressure p c is set in the engine room.
  • the apparatus solution in the prior art is on the Fig. 17 referenced, in which, as already mentioned above compressor according to the DE 102 29 659 A1 and according to the EP 1 154 160 A2 are shown.
  • the compressor according to the DE 102 29 659 A1 it is a system, which is controlled by the suction pressure p s , while the compressor according to the EP 1 154 160 A2 a regulation due to the differential pressure between suction pressure and high pressure (p d - p s ) takes place.
  • compressors according to the prior art Since operating speed of the compressor or operation of the vehicle almost constantly changes the speed (compressors according to the prior art are generally connected via a belt drive to the engine of the vehicle), in the compressors according to the prior art, permanent control interventions are necessary, ie a permanent variation of the engine room pressure p c is necessary (see also the above explanations).
  • the control process by a corresponding adjustment of the engine room pressure p c is slow in speed fluctuations and there is a strong overshoot, since the engine room comprises a relatively large volume.
  • the inertia of the control is due to the length of the controlled system, wherein the controlled system is as follows: In response to a change in the compressor speed n, the tilt angle of the swash plate changes, resulting in a change in the refrigerant mass flow, resulting in a change in the ratio of p d results to p s . After a detection of the ratio specified above, p c is readjusted as a function of the detected ratio.
  • Object of the present invention is to provide a method for controlling the refrigerant mass flow of a compressor, in which a largely constant refrigerant mass flow can be achieved even with speed fluctuations, with losses by regulating interventions between the pressure levels of high pressure p d and engine room pressure p c on the one hand or Engine room pressure p c and suction pressure p s (pressure in a suction chamber) on the other hand can be kept as low as possible by reducing the number of control interventions. Furthermore, care should be taken that one possible simple control valve configuration can be used, which ensures low cost. It is another object of the present invention to provide a compressor in which a method according to the invention is implemented.
  • the compressor for which the inventive method is designed generally has a drive shaft and an adjustable in its inclination to the drive shaft swash plate, which is arranged in a substantially defined by a housing of the compressor engine room and the same through their deflection angle with respect to the drive shaft defines the piston stroke s of the compressor.
  • the Sauggas In addition to the product of the compressor speed n, the Sauggas prevail p and Kolbenhubs s and the pressure p d on the high pressure side of the compressor and / or the pressure p s on the suction gas side of the compressor and / or a Regelventilstellish one Control valve (this is usually the energization of the coil of the valve), which is mounted between the high pressure side and the engine room in a compound of both chambers, held approximately constant.
  • a feedback feedback loop, usually a feedback controller on the control valve
  • the moment equilibrium is established between M sw and M k, ges (at least) for a deflection angle ⁇ gl for which ⁇ min ⁇ ⁇ gl ⁇ ⁇ max .
  • a predominance of the moment M sw (M sw > M k, ges ) can be produced, which is advantageous in particular when using the refrigerants R134a, R154a or the azeotropic mixtures or the hydrocarbons. This ensures an effective control characteristic.
  • the pressure p c in the engine room is varied in order to obtain a desired operating point, ie a desired refrigerant mass flow in the compressor. This ensures that desired operating points can be safely approached, while within one and the same operating point, ie at a certain desired refrigerant mass flow, the inventive method so automatically regulate engages that a changing engagement with respect to the engine room pressure p c is essentially not necessary ,
  • the suction pressure p s * and thus p is lowered so that the product of n, p and s is approximately constant.
  • the suction pressure p s in the sense of the present application, is the one prevailing in the cylinder space Suction pressure (which may well differ from the pressure p s in a suction chamber chamber upstream of the cylinders), while ⁇ represents the density of the refrigerant in the cylinder chamber or in the cylinder chambers.
  • the object is achieved by a compressor with the features of claim 10.
  • An essential point of the invention is that in a compressor, in particular axial piston compressor with a housing and a substantially arranged in the housing, driven via a drive shaft compressor unit for sucking and compressing a refrigerant with a likewise arranged in the housing swash plate, the moments M sw due the rotationally moving masses and M k, ges are due to the translationally moving masses to each other in a predetermined ratio, wherein the compressor comprises at least one intake-side inlet valve, which is configured so that it affects the entering into the cylinder chamber refrigerant mass flow speed-dependent so in that the ratio of the moments M sw and M k, ges on the one hand and the throttle power of the at least one intake valve on the other hand are related to each other in such a way that at least over parts of the possible speed range of the compressor, the refrigerant Ma stream, which is pumped into the system, is approximately constant.
  • the said parts of the speed range are preferably compressor speeds between 6000 and 9000 rpm, but in particular compressor speeds between 2500 and 7000 rpm.
  • the inlet valve or the intake valves can or can, for example, between a suction chamber (pressure p s ) and the cylinder chambers (pressure p s * ) can be arranged.
  • pressure p s suction chamber
  • pressure p s * cylinder chambers
  • Such a structural design of a compressor according to the invention ensures that control interventions for the engine room pressure p c are minimized, in particular in the case of speed jumps, since the refrigerant mass flow remains constant for a wide speed range without such control interventions.
  • the inlet valve which is responsible for the suction gas density p entering the cylinder, is a pressure-controlled lamella valve.
  • a slot-controlled valve can be arranged in the refrigerant circuit of a compressor according to the invention.
  • the at least one inlet valve, in particular lamella (or slot-controlled) valve preferably has a valve plate with through-hole (s) or through-flow bore (s) and a tongue-shaped suction lamella in particular.
  • Each cylinder can (may) be assigned one or more inlet valve (s), wherein additionally or alternatively, the corresponding suction lamellae can be integrated in a Sauglamellenplatine.
  • a slot-controlled valve may, for example, be a slot in the cylinder wall. In the field of lamellar valves, it may also be a construction in which the suction lamella is seated in the piston and the suction takes place under the piston. In all the above-described embodiments are constructionally easy to implement versions of a compressor according to the invention.
  • the end of one or each cylinder space associated with the inlet valve (s) may comprise an in particular radially extending annular enlargement which in particular limits the stroke of the suction lamella (s) and which is bevelled or flattened towards the attachment point of the suction lamella (s).
  • a refrigerant can find use in a compressor according to the invention preferably CO 2, thereby providing a more environmentally friendly use of the compressor (as well as environmentally friendly disposal thereof) is made possible.
  • the parameters listed below are preferably conceivable for using CO 2 as a refrigerant.
  • the ratio of piston diameter and through-bore in the valve plate (D / d) is about 1.5 to 5, especially 2.5 to 4, with a particularly preferred value being about 3.6.
  • the ratio of through-bore in the valve plate and lift of the suction fin (d / t) is about 2.5 to 8, especially 3.7 to 6.7, with a particularly preferred value being about 4 , 55 lies.
  • the ratio of piston stroke to the stroke of the suction plate in about 10 to 30, in particular 14 to 24, in which case a particularly preferred value is about 17.3. All the above-described values or ratios ensure that a compressor according to the invention has an optimum control behavior.
  • the design refers, as mentioned above, to the refrigerant R744 (CO 2 ), it being noted at this point that for other refrigerants, an adaptation of the parameter set is necessary and included in the spirit of the present invention.
  • R134a or R152a or azeotropic or azeotrope-like mixtures in particular mixtures of tetrafluoropropene and trifluoroiodomethane or hydrocarbons or mixtures of hydrocarbons, or halogenated or partially halogenated hydrocarbons, halogens, ethers, esters, alcohols or mixtures thereof and with one or more of the aforementioned substances used as a refrigerant.
  • a compressor according to the invention can have a piston diameter / piston stroke (D / s) ratio of about 0.6 to 1.6, in particular 0.875 to 1.4, in particular of about 1.14.
  • the ratio of piston diameter and passage throttle bore in the valve plate (D / d) is preferably in about 1.8 to 4, in particular 2.15 to 3.5, more preferably in about 2.8.
  • the ratio of passage throttle bore in the valve plate and stroke of the suction plate (d / t) in about 7 to 15, in particular 8.3 to 14.4, more preferably about 11.4.
  • the ratio of piston stroke to the lift of the suction lamella (s / t) may be approximately 15 to 40, in particular 20 to 36, and in particular 28.2.
  • the design described above refers to the refrigerants R143a, R152a, the azeotropic mixtures as well as hydrocarbons or mixtures of hydrocarbons.
  • the tilting behavior of the swash plate can be so effectively limiting effect that at high speeds of the same, especially at very high speeds or maximum speed, the angle of maximum deflection of the swash plate is smaller than the angle of maximum deflection ⁇ max at low speeds of compressor.
  • the geometry and dimensioning of all translationally moving parts such as axial piston, piston rod or sliding blocks or the like.
  • all rotationally moving parts such as swash plate, driver or the like.
  • the moment M k ges selected as a result of the translationally moving masses, in particular the piston, optionally including sliding blocks, piston rods or the like.
  • the angle of maximum deflection of the swash plate is smaller than the angle ⁇ max maximum deflection at lower speeds compressor.
  • a switching valve is arranged in a fluid connection arranged between the suction pressure level and the engine compartment and / or in a fluid connection arranged between the high pressure level and the engine compartment. This can be realized in a simple manner, a regulation of the engine room pressure p c .
  • a moment equilibrium between a torque M sw caused by rotationally moving masses and a moment M k, ges caused by translationally moving masses is determined for at least one deflection angle ⁇ gl of a swivel disk which is in the form of a swivel ring 1 (see. Fig. 2 ) is present.
  • Fig. 1 A clarification of the derivation of the two above-described moments is out Fig. 1 seen. This is a simplified derivation which is to be regarded as exemplary (in this context, simplifying is to be understood in the sense that in the model calculation the variables of interest for one slice are calculated) for the different moments.
  • ⁇ i ⁇ + 2 ⁇ ⁇ ⁇ i - 1 ⁇ 1 n
  • M sw J Y Z ⁇ ⁇ 2
  • a torque equilibrium (M k, ges ⁇ M sw ) for at least one deflection angle ⁇ gl of the swash plate or the swivel ring 1 is made, ie for a compressor in which the moment equilibrium feature of the inventive method is implemented in Fig. 2 shown.
  • a preferred embodiment of a compressor according to the invention comprises a housing, a cylinder block and a cylinder head.
  • pistons are mounted axially movable back and forth.
  • the compressor is driven by means of a belt pulley by means of a drive shaft 2.
  • the compressor described here is a variable piston-stroke compressor, the piston stroke being regulated by a pressure difference defined by the pressures p s * and p c ,
  • a pressure difference defined by the pressures p s * and p c
  • a swivel plate in the form of a swivel ring 1 more or less deflected or pivoted from its or its vertical position (see also Fig. 3b if the pressure difference is large, the tilt angle of the swing ring 1 is small, while if the pressure difference is small, the tilt angle is large).
  • the larger the resulting swing angle or deflection angle the larger the piston stroke. If the piston stroke is large, the mass flow is initially large.
  • the size of the corresponding pressure depends on the system control, ie the expansion device position.
  • a design with two springs is conceivable.
  • the sliding sleeve 3 can be stored both against the action of both springs, as well as with the action of a spring and against the action of the other spring.
  • the support member 5 is articulated both radially and (in a direction perpendicular to the drive shaft axis) perpendicular to the power transmission element 6, which means that the support member 5 is slidably mounted in a plane (and not only along an axis).
  • the support element 5 is formed in the shape of a cylinder bolt and has a groove 7, by means of which the support element 5 is in operative engagement with the force transmission element 6.
  • the support element 5 facing the end or is the support member 5 facing end portion of the power transmission element 6 in the form of a flat steel. This means that the said end region of the force transmission element 6 has an approximately rectangular peripheral contour. This approximately rectangular shaped end portion is engaged with the groove 7 of the support member 5 in engagement.
  • the advantage of the construction of the power transmission element 6 and the support member 5 and in particular their storage inside each other is that the flat steel does not have to build too high; the strength and rigidity (low deformation) is provided by the width of the bearing. In a central region, the strength of the force transmission element 6 increases while it is sleeve-shaped at its end facing the drive shaft 2. With the aid of the sleeve-shaped part 8 of the force transmission element 6 selbiges is mounted or fixed to the drive shaft 2. For a non-rotating connection of the drive shaft 2 with the sleeve-shaped part 8 of the force transmission element 6, a key 2a provides.
  • the power transmission element 6 is integrally formed and also einstoffig with the sleeve-shaped part 8.
  • the power transmission element 6 and the sleeve-shaped part 8 by two different components (possibly even of different materials) act.
  • the force transmission element 6 or the sleeve-shaped part 8 of the force transmission element 6 has two recesses in the form of grooves 9.
  • the power transmission element 6 and the sleeve-shaped part 8 can also be designed in one piece with the drive shaft 2. This can happen eg to trade a forging; a one-piece design is preferred for mass production.
  • the sleeve-shaped part 8 can be pushed in the assembled state of the swashplate mechanism under the spring 4. This means that the sleeve-shaped part 8 is placed over the drive shaft 2 and radially fixed by the spring 4 on the drive shaft 2.
  • the sliding sleeve 3 which has a recess 10 corresponding to the force transmission element 6, is slipped over the drive shaft 2 (sliding fit).
  • the sliding sleeve 3 also has two recesses in the form of holes 11.
  • the power transmission element 6 and the sliding sleeve 3 are secured by a groove nut (not shown) on the drive shaft 2, wherein the sliding sleeve 3 can reciprocate on the drive shaft 2 in the axial direction.
  • the sleeve-shaped part 8 of the power transmission element is fixed in rotation with the spring 4 on the drive shaft 2.
  • a plate spring 12 is further arranged on the drive shaft 2, which ensures that the compressor does not start at a minimum deflection angle of the pivot ring 1.
  • 2 stops in the form of stop plates 13, 14 are arranged on the drive shaft, which limit the deflection angle of the pivot ring.
  • the stop disc 13 serves as a stop for a minimum deflection angle
  • the stop plate 14 serves as a stop for a maximum deflection angle of the pivot ring 2.
  • On the back can also be provided a bearing seat for the main thrust bearing.
  • the support element 5 is mounted in a cylindrical recess in the form of a bore 15 in the pivot ring 1.
  • the bore 15 extends perpendicular to the drive shaft axis.
  • the power transmission element 6 is rotatably connected to the drive shaft 2 in the present preferred embodiment. It should also be noted at this point that the drive shaft 2 is not broken through the sleeve-shaped training or the sleeve-shaped part 8 of the power transmission element 6 and thus has corresponding stability.
  • the clear width of the bore of the pivot ring 1 is at least slightly larger than the corresponding extent of the force transmission element 6 (mountability).
  • the mechanism of support member 5 and power transmission element 6 is not intended to transmit the torque from the shaft to the swash plate in the form of the swivel ring 1.
  • the bearings between the support member 5 and the power transmission element 6, between the power transmission element 6 and drive shaft 2 and between the support member 5 and pivot ring 1 are not designed to transmit torque. It therefore eliminates a kind of driving function for the support member 5 and the power transmission element 6. This is deliberately chosen so for reasons of hysteresis, i. the tilting of the swivel ring 1 and the torque transmission are functionally decoupled from each other.
  • the mechanism of power transmission element 6 and support member 5 essentially receives the piston forces.
  • the torque in turn is transmitted from the drive shaft 2 to the swivel ring 1 by a tilting joint (realized by drive bolt 15a) provided on the drive shaft centerline.
  • the torque between the sliding sleeve 3 and the pivot ring 1 transmitting drive pin 15a are locked or secured to the pivot ring with snap rings 16a.
  • the swivel ring 1 has flats 17, which correspond to flats 18 on the sliding sleeve 3.
  • the sliding sleeve 3 to be omitted and for the torque transmission to take place in any desired form between the drive shaft and the swivel ring 1 (for example via flats on the drive shaft 2 and the swivel ring 1). It should be noted at this point that it is also within the scope of the present invention to couple the functions of torque transfer and gas power support.
  • Fig. 3a is a qualitative representation of the preferred interpretation of the moments according to the equations used (see. Fig. 1 ), wherein in addition to the rotational and translationally related moments (M sw and M k, ges ), the sum of the moments is shown. How one Fig. 3a can be found over a wide range of the tilt angle or deflection angle of the pivot ring 1 in about a moment equilibrium a, wherein it is a representation of the moments above the pivot angle ⁇ for any rotational speed n of the drive shaft 2 in the present figure.
  • Fig. 3b is for a fixed spring rate of the return spring 4 and for fixed pressure conditions on the high pressure and the suction side of 130 and 35 bar for different speeds n, the differential pressure prevailing between the engine room and suction side, each consideration for a compressor with moment equilibrium and with a prevailing over the entire speed range constant suction gas density p, applies. It should be noted at this point that this is a rather theoretical approach, since both the pressure p d and the pressure p s * at the piston for this calculation are assumed to be constant for each rotational speed n of the drive shaft 2. In practice, with increasing speed n, in particular the suction pressure p s * and thus the suction gas density p are lowered.
  • Fig. 3a the moment equilibrium M sw to M k, ges is qualitatively represented.
  • the moments M sw and M k, ges can also be adjusted by appropriate engine design so that in addition to an engine with neutral behavior as in Fig. 3a represented, an engine with alsregelndem behavior or an engine designed with abregelndem behavior.
  • the moments M sw and M k, ges or their relationship to each other would be (n) only provided accordingly, for CO 2 as the refrigerant course M sw ⁇ M k, ges or M sw > M k, ges is preferred.
  • the product of the compressor speed n, the Sauggas Together p of the sucked gas in the cylinders and the piston stroke s for different compressor speeds n at least for certain speed range in about kept constant, while a prevailing in the engine room pressure p c is also kept approximately constant (because, for example, the current flow of the control valve is kept constant).
  • the delivery volume, ie the mass flow can be kept approximately constant even without regulation of the engine room pressure p c .
  • the delivery volume per time V [cm 3 / s] V geo [cm 3 ] xn [1 / s], where V geo stands for the geometric delivery volume and n for the compressor speed.
  • V geo D 2 ⁇ / 4 xsx ⁇ , where ⁇ represents the number of pistons.
  • At least one intake-gas side inlet valve is arranged, which is configured such that the refrigerant mass flow entering the cylinder bores is speed-dependent (in particular on the Sauggas Stahl) influenced, that the ratio of the moments M sw and M k, ges on the one hand and the throttle performance the intake valve on the other hand in relation to each other in relation to each other that at least over parts of the speed range of the compressor, the refrigerant mass flow, which is conveyed in the system, is approximately constant.
  • the valves are used on the suction side conditionally as a throttle point and specifically in the context of the parameters M sw and M k, ges designed or tuned.
  • the deflection angle of the swivel ring 1 remains constant during speed jumps.
  • FIG Fig. 4a The throttling by the suction-side valves is shown schematically in FIG Fig. 4a shown, wherein the influence of throttling by a log-ph diagram in Fig. 4b is clarified (pressure reduction to p s , for the gas in the cylinder chamber).
  • the critical point (with CO 2 as refrigerant is a supercritical process) is designated KP.
  • KP The critical point
  • the real ratio is shown, which can vary depending on the operating point.
  • the selected representation is plotted for a fixed operating point. The illustration therefore applies to a constant tilt angle.
  • the influence of the throttling of the suction gas at a minimum tilt angle is very low and, for example, at a tilt angle of 0 ° does not exist, which is illustrated by the fact that meet in this area the curves of the individual speeds approximately.
  • Fig. 4c is the mass flow of the refrigerant over the speed shown qualitatively, which shows that the mass flow of the refrigerant is significantly reduced by the throttling effect at higher speeds than at low speeds.
  • the mass flow of the refrigerant is increased without loss from m2t to m3t. Due to the losses, however, the mass flow of the refrigerant is increased from a mass flow m2r to a mass flow m3r.
  • the swivel ring 1 is therefore a kind of internal controller (watt controller).
  • Undercompensation means in this context that at a doubling of the rotational speed, the geometric displacement or the tilt angle or the stroke of the piston is so automatically changed that the mass flow of the refrigerant is slightly reduced compared to the starting position. A corrective control intervention becomes necessary.
  • overcompensation in this context means that at a doubling of the rotational speed, the geometric displacement or the tilt angle or the stroke of the piston is automatically changed such that the mass flow of the refrigerant is slightly increased compared to the starting position. A corrective control intervention becomes necessary, as in the case of undercompensation.
  • FIG. 6 Indicator diagrams for two operating points are shown to show the influence of the valve losses as a function of the rotational speed n of the drive shaft 2. While at a speed of 800 rpm, the average pressure loss is about 0.5 bar, the pressure loss is at the same valve configuration at 3000 rev / min on average about 3 bar. This behavior can be influenced by appropriate dimensioning of the suction-side valves within certain limits.
  • the dimensioning of the suction-side valves and the compressor geometry is in Fig. 7 described.
  • the dimensioning of the parameters refers to the application of the refrigerant R744 (CO 2 ).
  • the dimensioning of compressors which use refrigerant R134a / R152a varies considerably; Here, the vote of the moment equilibrium or the moments M sw and M k, ges should look significantly different with respect to the valve geometry. In R134a / R152a, the pressure losses are relatively lower, resulting in that the moments M sw must be chosen to be greater than M k, ges (overcompensation of the moments) in order to achieve compensation in the range of the mass flow of the refrigerant.
  • the compressor has (see. Fig. 7 ) On the inlet side for the suction gas in the cylinder chamber, a valve plate 19 with a suction lamella 20 mounted thereunder.
  • the suction lamella 20 is tongue-shaped and serves to control the Sauggaseinlasses.
  • the suction lamella 20 closes a through-flow bore 21, while the suction lamella 20 moves downwards during aspiration of the suction gas (due to the negative pressure prevailing in the cylinder) by a stroke t (indicated by arrows 22) and to be sucked in Refrigerant or the suction gas through the passage throttle bore 21 inlet into the cylinder granted.
  • the passage throttle bore 21 has a diameter d. Due to the geometry of the inlet valve, ie in particular due to the diameter d of the passage throttle bore 21 or in particular due to the sum of the diameter d of the passage throttle bore 21 and the stroke t of the suction plate 20 and the compressor geometry over a wide work areas of the compressor according to the invention to a desired lowering the suction pressure p s .
  • the number of pistons N is 5 to 9; the stroke t of the suction lamella 20 is between 0.9 and 1.2 mm, while the valve plate 19 has a bore (through-flow bore 21) whose diameter d is between 4 and 6 mm.
  • the values for the piston diameter D are approximately 15 to 19 mm and the piston stroke s is approximately 17 to 22 mm.
  • the maximum stroke volume per cylinder V is 3 ccm to 6 ccm.
  • 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 21 in the valve plate 19 of about 2.5 to 4, a ratio of passage throttle bore 21 in the valve plate 19 and stroke t of the suction plate of about 3.7 to 6.7 and a ratio of piston stroke s to the stroke t of the suction plate of about 14 to 24th
  • the passage throttle bore 21 is used on the suction side as a throttle point and designed specifically in conjunction with the other parameters controlling the compressor.
  • the inflowing gas flows through a suction chamber, which is mounted in the cylinder head, with the pressure P s 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 s * adjusts, which ensures an optimal control behavior of the compressor.
  • Fig. 8 finally a speed jump from 2000 rpm to 6000 rpm is shown; the curves represent the pressure at the suction gas side, the mass flow of the refrigerant, the speed and the pressure at the high pressure side.
  • the mass flow of the refrigerant and the pressures in the engine room at the suction gas side of the compressor and the pressure side of the compressor remain substantially unchanged.
  • According to the invention has been achieved by a vote of the moments M sw and M k, ges in connection with the suction valves that prevails this behavior.
  • the ideal range for the design is, as already mentioned, the average speed range, so that for the above sizes short-term changes (mass flow of the refrigerant and the pressures in the engine room on the suction side of the compressor and the pressure side of the compressor) are compensated josschreib.
  • the mass flow for a corresponding speed jump can be measured relatively easily.
  • the design of the parameters of the suction valves and the parameters M sw and M k, ges can be understood by measuring and weighing.
  • a measurement of the piston stroke can be done by attaching a magnet to the piston in a simple manner, since the magnet can be detected via a sensor on the housing.
  • the mass flow m is detectable before or after the compressor mass flow meters.
  • By means of a tachometer and the compressor speed n can be determined in a simple manner.
  • FIG. 9 Another preferred embodiment of the swash plate mechanism of a compressor according to the invention is in Fig. 9 shown.
  • this mechanism for CO 2 as a refrigerant as well as for the already mentioned refrigerant R134a or R152a or the azeotropic or azeotrope-like Mixtures, in particular mixtures of tetrafluoropropene and trifluoroiodomethane or the refrigerants of one or more hydrocarbons (hydrocarbons or mixtures thereof) can be used.
  • the swiveling ring is designed in such a way that for at least one deflection angle ⁇ of the swivel ring 1, a moment equilibrium M k, ges ⁇ M sw occurs, while when the other refrigerants are used, a predominance of the moment M sw in relation to FIG Moment Mk, ges ( Msw ⁇ Mk, ges ) is advantageous for at least a portion of the possible deflection angle of the pivot ring.
  • the swashplate mechanism according to Fig. 9 is largely analogous to that according to Fig. 2 Therefore, in particular the differences to the mechanism according to Fig. 2 exposed.
  • swash plate mechanism also has the one according to Fig. 9 a swivel ring 1, a drive shaft 2, a sliding sleeve 3, and a support member 5 and a power transmission element 6.
  • a swivel ring 1 a swivel ring 1
  • a drive shaft 2 a sliding sleeve 3
  • a support member 5 a support member 5 and a power transmission element 6.
  • Fig. 2 Analogous to the in Fig. 2 shown swash plate mechanism also has the one according to Fig. 9 a swivel ring 1, a drive shaft 2, a sliding sleeve 3, and a support member 5 and a power transmission element 6.
  • the support element 5 does not have a groove-shaped or pocket-shaped recess, but has a roughly rectangular recess 23, which extends in the radial direction through the entire support element 5.
  • the power transmission element 6 is approximately cylindrical in shape and has at its radially outer end (ie, at its end facing the pivot ring 1) a flat steel-like design. This engages in the recess 23 and thus forms the articulation of the power transmission element 6 to the support element. 5
  • the cross section of the recess 23 increases radially outward, ie, the cross section widens in said radially outer region in an approximately V-shaped, while in a radially inner region (from the Fig. 9 not apparent) is approximately constant.
  • the radially outer, frontal edges of the power transmission element 6 are arranged for each tilt angle of the pivot ring 1 in the radially outer region of the recess 23. This means that the edges for each tilt angle of the swivel ring 1 over the region of the approximately constant Aussparungsqueritess and thus avoid tilting with the support member at each operating point of the compressor.
  • the power transmission element 6 is with its drive shaft 2 facing the end in the drive shaft 2, ie in a corresponding recess 24 of the drive shaft is pressed.
  • a semicircular or groove-shaped recess 25 which is part of a fluid connection between the engine room and the engine chamber of the compressor and the suction gas side of the compressor. This fluid connection serves to regulate the pressure in the engine chamber.
  • the corresponding fluid connection extends through the drive shaft and connects the engine room, in which the pressure prevails p c , with the suction side of the compressor and / or the high pressure side of the compressor.
  • a piston 26 is shown, which is articulated by means of two sliding blocks 27 to the pivot ring 1.
  • Fig. 10 is analogous to Fig. 3b a diagram is shown, which represents the differential pressure or the pressure in the engine room as a function of the geometric displacement, ie, where appropriate, the lift or tilt angle for a compressor with the above-specified refrigerant (R134a, etc.).
  • R134a refrigerant
  • a preferred for the said engine engine design is adjustable, which creates an engine, wherein in the range of the possible tilt angle of the pivot ring 1 M sw ⁇ M k, ges holds. It can be seen that the control curves intersect in exactly one point.
  • Fig. 11a and 11b Analogous to Fig. 3a are in the Fig. 11a and 11b in each case the moments M k, ges and M sw as well as the sum of M k, ges and M sw (M ges ) are represented for two different parameter sets over the tilt angle of the swash plate. While in Fig. 11b a more characteristic of the refrigerant CO 2 characteristic is shown in Fig. 11a a characteristic more conducive to the other refrigerants listed above (R134a, etc.) ( Msw ⁇ Mk, ges ).
  • n (p) the number of pistons
  • R the distance of the piston axis to the shaft axis
  • n the shaft speed
  • mk the mass of a piston including sliding block
  • mk ges the mass of all pistons including sliding block
  • L the position of the pivot thread
  • msw the mass of the swash plate
  • ra the outer radius of the swash plate
  • ri represents the inner radius of the swash plate
  • h the height of the swash plate.
  • the Fig. 12a to 12c correspond to the Fig. 4a to 4c , where the Fig. 4a to 4c for CO 2 as refrigerant have validity, while the 12a to 12c Valid for the other refrigerants (R134a etc.).
  • the log-ph diagram according to Fig. 12b is adapted to the said refrigerant, since in contrast to CO 2 as a refrigerant in the mentioned other refrigerants, a liquefaction (wet steam area) takes place.
  • Fig. 12d the individual states of the refrigerant associated with a refrigeration circuit shown schematically.
  • Fig. 13 sets the analogue to Fig. 5 for the mentioned refrigerant (R134a etc.). Shown is the modified tilt characteristic with respect to Fig. 10 due to the effect on the suction and intake valves. While the remarks to Fig. 5 Keep validity, it should be noted that at higher compressor speeds of the swivel ring can not be adjusted to a maximum tilt angle. This is an important safety feature of the inventive compressor concept.
  • Fig. 14 Analogous to the illustrations in Fig. 6 is in Fig. 14 a representation of four indicator diagrams for the mentioned refrigerant (except CO 2 ) is given, it being noted that the representations qualitatively in about all enumerated refrigerant (including CO 2 ) can be used.
  • piston diameter and piston stroke (D / s) of about 0.6 to 1.6, in particular 0.875 to 1.4, more particularly of about 1.14.
  • the ratio of piston diameter and fürgangsdrosselbohrung in the valve plate (D / d) is about 1.8 to 4, in particular 2.15 to 3.5, more particularly in about 2.8.
  • the ratio of passage throttle bore in the valve plate and stroke of the suction plate (d / t) is in about 7 to 15, in particular 8.3 to 14.4, more particularly in about 11.4, and the ratio of piston stroke to the stroke of the suction plate ( s / t) is about 15 to 40, in particular 20 to 36, furthermore in particular about 28.2. It should be noted at this point that the parameters given above are valid for the refrigerants R134a, R152a, etc.
  • Fig. 15 show the parameters to be used for the design of the moment equilibria. In the various columns, first a preferred parameter set is given; Furthermore, the parameters are still determined area by area. These parameters are also to be preferred for the refrigerants R134a, R152 etc.
  • a simple switching valve can be used, which can influence the gas flow from the high-pressure side into the engine room.
  • the switching valve can intervene when another operating point is to be set.
  • An intervention on the control valve by a so-called feedback as in the prior art is not necessary.
  • the control valve which regulates the gas flow from the pressure side of the compressor in the engine room of the compressor, thus no additional signal must be supplied, as is known in the prior art.
  • additional signals e.g. the change in the mass flow of the refrigerant, the change of a pressure difference, the change of the suction pressure, etc.
  • the self-regulation can compensate for variations in the refrigerant mass flow due to the rotational speed. It should be noted at this point that it is essential that not only the mass flow can be kept substantially constant, but at the same time the pressure layers on the pressure side and the suction side of the compressor.
  • the solenoid of the control valve actuates the control valve only when a new operating point is to be set.
  • a so-called switching valve is compared to the prior art thus characterized in that the feedback range can be omitted. Such a switching valve is significantly cheaper than the valves used in the prior art.
  • Such a simple valve used in a compressor according to the invention is preferably a valve of the type used for today's ABS or ESP valves.
  • inventive scheme works much faster than the previous scheme.
  • the variables to be controlled are regulated approximately at the same time as the increase in the rotational speed; according to the prior art, this happens with a time delay, since first a feedback variable must be able to be picked up, which is supplied or assigned to the control valve.

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Claims (27)

  1. Procédé de régulation du débit massique de frigorigène d'un compresseur, en particulier d'un compresseur à pistons axiaux, plus particulièrement pour des climatisations de véhicule automobile, qui présente un arbre d'entraînement (2) et un plateau pivotant (1) dont l'inclinaison par rapport à l'arbre d'entraînement est réglable, plateau qui est disposé dans une chambre de mécanisme d'entraînement du compresseur essentiellement définie par un carter du compresseur et qui définit la course des pistons du compresseur par son angle de déviation par rapport à l'arbre d'entraînement (2),
    sachant qu'on produit pour au moins un angle de déviation du plateau pivotant (1) un équilibre de couples approximatif entre un couple Msw dû aux masses déplacées en rotation et un couple Mk,ges dû aux masses déplacées en translation (Msw = Mk,ges), et/ou une prépondérance du couple Msw par rapport au couple Mk,ges pour au moins une plage des angles de déviation possibles du plateau pivotant (MSW ≥ Mk,ges),
    caractérisé en ce que le produit du régime n du compresseur, de la densité p du gaz d'aspiration et de la course s des pistons est automatiquement maintenu approximativement constant pour différents régimes n de compresseur au moins dans certaines plages, en particulier pour des régimes n de compresseur compris entre 600 et 9000 tr/min, notamment pour des régimes de compresseur compris entre 2500 et 7000 tr/min, tandis qu'une pression pc régnant dans la chambre de mécanisme d'entraînement est également maintenue environ constante (ρ x s x n = constante),
    sachant qu'on réalise une prépondérance du couple Msw (MSW > Mk,ges) pour tous les angles de déviation.
  2. Procédé selon la revendication 1, caractérisé en ce qu'en plus du produit du régime n du compresseur, de la densité p du gaz d'aspiration et de la course - s des pistons, on maintient également approximativement constantes la pression pd sur un côté haute pression du compresseur, et/ou la pression ps sur un côté de gaz d'aspiration du compresseur, et/ou une grandeur réglée d'une soupape de régulation entre le côté haute pression et la chambre de mécanisme d'entraînement.
  3. Procédé selon la revendication 1 ou 2, caractérisé en ce qu'on réalise l'équilibre de couples entre Msw et Mk,ges au moins pour αgl (αmax - αmin)/2, αmax désignant l'angle de déviation maximal du plateau pivotant (1) et αmin l'angle de déviation minimal du plateau pivotant (1)
  4. Procédé selon l'une des revendications précédentes, caractérisé en ce qu'on réalise l'équilibre de couples entre Msw et Mk,ges au moins pour un angle de déviation agl qui satisfait à la relation : αmin ≤ αgl ≤ αmax.
  5. Procédé selon l'une des revendications précédentes, caractérisé en ce qu'on réalise l'équilibre de couples entre Msw et Mk,ges au moins pour un angle de déviation agl = αmax.
  6. Procédé selon l'une des revendications précédentes, caractérisé en ce qu'on réalise l'équilibre de couples entre Msw et Mk,ges au moins pour un angle de déviation (fictif) αgl ≥ αmax.
  7. Procédé selon l'une des revendications précédentes, caractérisé en ce qu'on fait varier la pression pc dans la chambre de mécanisme d'entraînement afin d'obtenir un point de fonctionnement souhaité, c'est-à-dire un débit massique souhaité de frigorigène dans le compresseur.
  8. Procédé selon l'une des revendications précédentes, caractérisé en ce que, en cas d'augmentation du régime n, on abaisse la pression d'aspiration ps, et donc p, de telle sorte que le produit de n, p et s est approximativement constant.
  9. Compresseur, en particulier compresseur à pistons axiaux, plus particulièrement compresseur pour la climatisation d'un véhicule automobile, avec un boîtier, et avec une unité de compression pour aspirer et comprimer un frigorigène, disposée pour l'essentiel dans le boîtier, entraînée au moyen d'un arbre d'entraînement (2) et pourvue d'un plateau pivotant (1) lui aussi dans le boîtier,
    caractérisé en ce que le couple MSW dû aux masses déplacées en rotation et le couple Mk,ges dû aux masses déplacées en translation se trouvent dans un rapport prédéfini l'un par rapport à l'autre,
    et en ce que le compresseur comprend au moins une soupape d'admission disposée côté gaz d'aspiration, qui est configurée de telle sorte qu'elle influe, en fonction du régime, sur le débit massique de frigorigène parvenant dans la chambre de cylindre de telle sorte que le rapport des couples MSW et Mk,ges d'une part et de l'étranglement produit par la soupape d'admission au moins unique d'autre part se trouvent dans une relation mutuelle telle que, au moins sur une partie de la plage de régimes, en particulier pour des régimes n de compresseur compris entre 600 et 9000 tr/min, notamment pour des régimes de compresseur compris entre 2500 et 7000 tr/min, le débit massique de frigorigène véhiculé dans le système est approximativement constant,
    sachant qu'on peut réaliser une prépondérance du couple Msw (MSW > Mk,ges) pour tous les angles de déviation.
  10. Compresseur selon la revendication 9, caractérisé en ce qu'au moins une soupape d'admission est une soupape à lamelles (19, 20, 21) pressostatique.
  11. Compresseur selon la revendication 9 ou 10, caractérisé en ce qu'au moins une soupape d'admission est une soupape commandée par fente.
  12. Compresseur selon l'une des revendications 9 à 11, caractérisé en ce que la soupape d'admission, notamment la soupape à lamelles, comprend un disque de soupape (19) doté d'un orifice d'étranglement traversant (21), et une lamelle d'aspiration (20) notamment en forme de languette.
  13. Compresseur selon l'une des revendications 9 à 12,
    avec un bloc-cylindres et au moins un et notamment 5 à 9 pistons qui sont mobiles axialement en va-et-vient dans des alésages prévus dans le bloc-cylindres,
    caractérisé en ce qu'une soupape d'admission est associée à chaque cylindre, et/ou les lamelles d'aspiration correspondantes (20) sont intégrées dans une platine de lamelles d'aspiration.
  14. Compresseur selon l'une des revendications 9 à 13, caractérisé en ce que l'extrémité de la ou de chaque chambre de cylindre qui est associée à la soupape d'admission comprend un élargissement annulaire s'étendant radialement, qui limite notamment la course de la/des lamelles d'aspiration (20) et qui est chanfreiné ou aplati en direction du point de fixation de la/des lamelles d'aspiration (20).
  15. Compresseur selon l'une des revendications 9 à 14, caractérisé en ce qu'on utilise comme frigorigène du CO2.
  16. Compresseur selon l'une des revendications 9 à 15, notamment selon la revendication 15, si selon l'une des revendications 8 à 12 ou 14, avec en outre un bloc-cylindres et au moins un et notamment 5 à 9 pistons qui sont mobiles axialement en va-et-vient dans des alésages prévus dans le bloc-cylindres,
    caractérisé en ce que le rapport du diamètre de piston à la course de piston (D/s) est d'environ 0,4 à 1,5, en particulier de 0,65 à 1,1, plus particulièrement d'environ 0,95.
  17. Compresseur selon l'une des revendications 9 à 16, notamment selon la revendication 15 ou 16, caractérisé en ce que le rapport du diamètre de piston à l'orifice d'étranglement traversant du disque de soupape (D/d) est d'environ 1,5 à 5, en particulier de 2,5 à 4, plus particulièrement d'environ 3,6.
  18. Compresseur selon l'une des revendications 9 à 17, notamment selon l'une des revendications 15 à 17, caractérisé en ce que le rapport de l'orifice d'étranglement traversant du disque de soupape à la course de la lamelle d'aspiration (d/t) est d'environ 2,5 à 8, en particulier de 3,7 à 6,7, plus particulièrement d'environ 4,55.
  19. Compresseur selon l'une des revendications 9 à 18, notamment selon l'une des revendications 15 à 18, caractérisé en ce que le rapport de la course de piston à la course de la lamelle d'aspiration (s/t) est d'environ 10 à 30, en particulier de 14 à 24, plus particulièrement d'environ 17,3.
  20. Compresseur selon l'une des revendications 9 à 14, caractérisé en ce qu'on utilise comme frigorigène du R134a ou du R152a ou des mélanges azéotropiques ou de comportement analogue aux azéotropes, notamment des mélanges de tétrafluoropropène et de trifluoroiodométhane ou d'hydrocarbure(s), ou des mélanges d'hydrocarbures, ou des hydrocarbures halogénés ou partiellement halogénés, des halogènes, des éthers, des esters, des alcools ou des mélanges de ceux-ci et d'une ou plusieurs des substances précitées.
  21. Compresseur selon l'une des revendications 9 à 14 ou 20, notamment selon la revendication 20,
    si selon l'une des revendications 8 à 12 ou 14 ou 20, avec en outre un bloc-cylindres et au moins un et notamment 5 à 9 pistons qui sont mobiles axialement en va-et-vient dans des alésages prévus dans le bloc-cylindres, caractérisé en ce que le rapport du diamètre de piston à la course de piston (D/s) est d'environ 0,6 à 1,6, en particulier de 0,875 à 1,4, plus particulièrement d'environ 1,14.
  22. Compresseur selon l'une des revendications 9 à 14 ou 20 à 21, notamment selon la revendication 20 ou 21, caractérisé en ce que le rapport du diamètre de piston à l'orifice d'étranglement traversant du disque de soupape (D/d) est d'environ 1,8 à 4, en particulier de 2,15 à 3,5, plus particulièrement d'environ 2,8.
  23. Compresseur selon l'une des revendications 9 à 14 ou 20 à 22, notamment selon l'une des revendications 20 à 22, caractérisé en ce que le rapport de l'orifice d'étranglement traversant du disque de soupape à la course de la lamelle d'aspiration (d/t) est d'environ 7 à 15, en particulier de 8,3 à 14,4, plus particulièrement d'environ 11,4.
  24. Compresseur selon l'une des revendications 9 à 14 ou 20 à 23, notamment selon l'une des revendications 20 à 23, caractérisé en ce que le rapport de la course de piston à la course de la lamelle d'aspiration (s/t) est d'environ 15 à 40, en particulier de 20 à 36, plus particulièrement d'environ 28,2.
  25. Compresseur selon l'une des revendications 9 à 24, caractérisé en ce que le comportement de basculement du plateau pivotant (1) agit de façon automatiquement limitatrice de telle sorte qu'à des régimes élevés du compresseur, notamment à des régimes très élevés ou au régime maximal, l'angle de déviation maximale du plateau pivotant (1) est inférieur à l'angle de déviation maximale αmax à des régimes inférieurs du compresseur.
  26. Compresseur selon l'une des revendications 9 à 25, caractérisé en ce que la géométrie et le dimensionnement, d'une part de tous les éléments déplacés en translation, tels que pistons axiaux, tiges de piston ou coulisseaux ou analogues, et d'autre part de tous les éléments déplacés en rotation, tels que plateau pivotant (1), entraîneurs ou analogues, sont tels que, pour des angles prédéfinis de basculement du plateau pivotant (1), notamment entre un angle de basculement minimal prédéfini et un angle de basculement maximal prédéfini, le couple Mk,ges dû aux masses déplacées en translation, notamment aux pistons, le cas échéant y compris les coulisseaux, tiges de piston ou analogues, est choisi inférieur au couple Msw dû au couple de déviation, c'est-à-dire au couple dû à l'inertie de masse du plateau pivotant, dans une mesure telle qu'à des régimes élevés du compresseur, notamment à des régimes très élevés ou à un régime maximal, l'angle de déviation maximale αmax du plateau pivotant (1) est inférieur à l'angle de déviation maximale à des régimes inférieurs du compresseur.
  27. Compresseur selon l'une des revendications 9 à 26,
    qui présente une liaison fluidique entre un côté haute pression et une chambre de mécanisme d'entraînement et/ou entre un côté de pression d'aspiration et la chambre de mécanisme d'entraînement,
    caractérisé en ce qu'une soupape de commande est disposée dans les ou dans au moins une des liaisons fluidiques.
EP07725250A 2006-05-23 2007-05-15 Procédé de régulation du débit massique de fluide frigorigène d'un compresseur Not-in-force EP2021631B1 (fr)

Applications Claiming Priority (4)

Application Number Priority Date Filing Date Title
DE102006024533 2006-05-23
DE102006029875A DE102006029875A1 (de) 2006-05-23 2006-06-28 Verfahren zum Regeln des Kältemittel-Massenstroms eines Verdichters
DE102006029874A DE102006029874A1 (de) 2006-05-23 2006-06-28 Verfahren zum Regeln des Kältemittel-Massenstroms eines Verdichters
PCT/EP2007/004334 WO2007134760A1 (fr) 2006-05-23 2007-05-15 Procédé de régulation du débit massique de fluide frigorigène d'un compresseur

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EP2021631A1 EP2021631A1 (fr) 2009-02-11
EP2021631B1 true EP2021631B1 (fr) 2010-05-12

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DE19616961C2 (de) * 1996-04-27 2002-11-07 Daimler Chrysler Ag Hubkolbenmaschine mit Taumelscheibengetriebe
DE19839914A1 (de) * 1998-09-02 2000-03-09 Luk Fahrzeug Hydraulik Axialkolbenmaschine
DE10329393A1 (de) * 2003-06-17 2005-01-05 Zexel Valeo Compressor Europe Gmbh Axialkolbenverdichter, insbesondere Kompressor für de Klimaanlage eines Kraftfahtzeuges
DE102004040042A1 (de) * 2004-08-18 2006-02-23 Zexel Valeo Compressor Europe Gmbh Axialkolbenverdichter

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ATE467761T1 (de) 2010-05-15
WO2007134760A1 (fr) 2007-11-29
DE502007003747D1 (de) 2010-06-24
EP2021631A1 (fr) 2009-02-11

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