DE102006029875A1 - Method for controlling the coolant mass stream of an axial piston compressor comprises guiding a pivoting disk in equilibrium between a torque produced by rotating masses and a torque produced by translating masses - Google Patents

Abstract

Method for controlling the coolant mass stream of an axial piston compressor comprises guiding a pivoting disk in equilibrium between a torque produced by rotating masses and a torque produced by translating masses and keeping constant the product of the compressor rotation, a suction gas density and piston lift for different revolutions whilst a pressure prevailing in the propellant chamber is kept constant. An independent claim is also included for an axial piston compressor. Preferred Features: The equilibrium between the torques is produced for an ejection angle.

Description

The The present invention relates to a method of controlling a refrigerant mass flow a compressor according to claim 1 and a compressor according to the preamble of claim 9.

Compressors for automotive air conditioning systems and method for controlling the same are known from the prior art. There are increasingly axial piston compressor use in automotive air conditioning systems. 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. Due to the ratio of the prevailing in the engine room pressure p _c and the pressure on a high pressure side of the compressor p _d or the pressure on a suction gas side of the compressor p _s , a desired deflection angle of the swash plate (corresponding to a certain stroke height of the piston) by a change in pressure p _{c are produced} in the engine room, which ensures a desired mass flow of refrigerant. The disadvantage of this, however, is that especially at changing rotational speeds (the speed changes almost constantly, since the compressor is connected via a belt drive with the engine) very many control interventions for the prevailing pressure in the engine room p _{c are} necessary.

That is why in the EP 0 809 027 A1 expressed the desire that the flow rate of a compressor should be compensated by the dynamic behavior of the engine of the same, so that the flow rate can be kept constant. Furthermore, it is stated in the said application that for constant control of the flow rate at varying rotational speeds, the restoring torque of a swash plate can be used, which counteracts their inclination due to dynamic forces on the co-rotating disc part.

On the EP 0 809 027 A1 constructively reveals the DE 198 39 914 A1 Measures, such as such a control behavior (an at least partial compensation of the flow rate) can be achieved. It is proposed to dimension the component mass of the swash plate with respect to the translationally moved masses so that the centrifugal forces of the swash plate affect the control behavior of the same. This is to be achieved essentially by the fact that the rotating mass of the drive pulley or the pivotable portion of the drive pulley is greater than the common mass of all pistons, so that the centrifugal forces occurring when turning the drive pulley sufficient to deliberately counteract the pivotal movement of the drive pulley and Thus, to influence the piston stroke and thus the flow rate, in particular to reduce or limit.

In the case of the applicant DE 103 29 393 It is explained why the component mass should not be the preferred parameter to influence the control performance of the engine due to speed variations as desired. It is further stated that the desired control behavior of the compressor can not be achieved with the component mass of the swash plate in relation to the translationally moving masses, but only taking into account the moment of inertia of the swash plate, which depends more on the geometry of the same as the component mass. A core idea of DE 103 29 393 is to compensate for speed fluctuations or changes in speed, the torque due to translationally moving masses directly by the moment due to rotating masses or overcompensate.

In the also attributed to the applicant DE 103 47 709 A1 It is proposed to tune the effective moments due to the inertial forces and the moments due to the Deviationsmomente so that the Schrägscheibenkippwinkel largely does not change at varying speeds. In the further is from the DE 103 47 709 A1 It is known that such a torque distribution results in an engine behavior which optimally regulates the mass flow of a refrigerant compressor, but an optimum refrigerant mass flow can only be achieved for a limited speed range of the compressor.

Further explanations for the control of a compressor are, for example, the DE 195 14 748 A1 refer to. A compressor described therein provides (as already explained above) a suction gas pressure level p _s and a high-pressure level p _d during operation. This is done by (usually) a control valve, through which the operating point is set. Likewise, the air conditioner about these pressure levels. Responsible for this is, for example, an expansion device (which regulates the circulation, pressure separation), which in turn reacts to changes in the operating state of the compressor and, where appropriate, regulating attacks. In the compressor engine room, for example, a pressure p _{c is} set by control valves on the compressor, which is between the Sauggasdruckniveau 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 between p _d and p _c , volume flow 2 between p _c and p _s ) the individual chambers or pressure positions. The model described here is simplified and should be considered as an example.

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).

In addition to a reduced performance of the vehicle by the high number of power-consuming control interventions is to be considered and disadvantageous that 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: On a change in the compaction terdrehzahl n down the tilt angle of the swash plate changes, which has a change in the refrigerant mass flow result, resulting in a change in the ratio of p _d results to p _s . After detection of the above-described ratio, pC is readjusted depending on the detected ratio.

By the engine concepts and the method for controlling a compressor according to the EP 0 809 027 , of the DE 198 39 914 and the DE 103 29 393 Compressors are known in which an increase in the mass flow, for example as a result of an increase in speed is countered by the fact that the tilt angle of the swash plate decreases. The concept is based on achieving this by means of a corresponding design of the swashplate, whereby an overcompensation of the txanslatorisch moving masses (piston and sliding blocks) is provided. The problem with this type of design or regulation is that the mass and the geometry of the swash plate must be comparatively large and that the moment equilibrium or overcompensation of translationally moving masses not over the entire speed range of a compressor, but only for a relatively narrow Speed corridor is possible.

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 the simplest control valve configuration can be used, which ensures low costs. It is another object of the present invention to provide a compressor in which a method according to the invention is implemented.

These Task is achieved by a method having the features of claim 1, wherein advantageous developments and details of the invention in the dependent claims are described.

In detail, the object is achieved by a method for controlling the refrigerant mass flow of a compressor, in particular an axial piston compressor and further in particular a compressor for motor vehicle air conditioning systems, which may comprise CO ₂ as a refrigerant, in which approximately a moment equilibrium between a by rotary moving masses induced moment M _SW and a due translationally moved masses moment M _{k, ges} for at least one deflection angle α _g1 of the swash plate is brought about, and in which n is the product of the compressor speed, the suction gas ρ (the suction gas, which is an in-cylinder with pressure p _{s *} flows and the pressure p _{s *} may be different from p _s, since by constrictions, or the like, a pressure reduction between a Sauggaskammer and the cylinder space occurs) and the piston stroke s for different compressor speeds n, at least for. certain speed ranges, in particular for compressor speeds n between 600 and 9000 U / min, and further in particular for compressor speeds between 2500 and 7000 U / min, automatically kept approximately constant, while the prevailing in the engine room pressure p _{c is} also kept approximately constant. It should be noted that 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.

In a preferred embodiment of the method according to the invention, in addition to the product of the compressor speed n, the Sauggasdichte ρ and the piston stroke 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 Regelventilstellgröße 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. By the method features described above it is ensured that only a small number of control actions for the engine room pressure p _{c is} necessary. In particular, by the features of claim 1 ensures that a feedback (feedback loop, usually a feedback controller on the control valve) can be dispensed with, which is necessary for the control of compressors or in control method according to the prior art.

Preferably, the moment equilibrium between M _SW and M _{k, ges is established} at least for α _g1 = (α _max -α _min ) / 2, it being noted that α _max denotes the maximum deflection angle of the swash plate and α _min the minimum deflection angle of the swash plate. Alternatively or additionally, the moment equilibrium between M _SW and M _{k, ges} produced (at least) for a deflection angle α _g1, for which applies α _min ≤ α ≤ α _{g1 max.} Additionally or alternatively, the torque balance between M _SW and M _{k, ges} (at least) for a deflection angle α _g1 - α _max and / or for a (fictitious) deflection angle α _g1 ≥ α _max can be produced. This ensures an effective control characteristic.

In a further preferred embodiment, 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 ,

In a further preferred embodiment of a method according to the invention, with an increase in the compressor speed n, the suction pressure p _{s *} and thus ρ are lowered in such a way that the product of n, ρ and s is approximately constant. For the purposes of the present application, the suction pressure p _{s *} is the suction pressure prevailing in the cylinder space (which may well differ from the pressure p _s in a suction gas chamber upstream of the cylinders), while ρ is the density of the refrigerant in the cylinder space or in the cylinder spaces represents.

Regarding the device-technical aspect, the task is through a compressor with the features of claim 9 solved.

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 throttling power of the at least one inlet 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 intake valve or intake valves may be arranged, for example, between a suction gas chamber (pressure p _s ) and the cylinder chambers (pressure p _{s *} ). In addition to the mere existence of the intake valves and their dimensioning, which can lead to certain desired losses in the refrigerant mass flow, as well as the Abstim mung refrigerant mass flow loss-causing measures lead each other to the desired control characteristic or contribute. 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.

In a preferred embodiment is the inlet valve, which for the suction gas density ρ reaching the cylinder is responsible pressure-controlled flap valve. Alternatively or additionally, too a slot-controlled valve in the refrigerant circuit of a compressor according to the invention be arranged. The at least one inlet valve, in particular lamellar (or slot-controlled) valve, preferably has a valve plate with through hole (s) or through throttle bore (s) and a tongue-shaped in particular Suction lamella on. Each cylinder can have one or more inlet valves (s) be assigned, in addition or alternatively, the corresponding suction lamellae in one Suction lamella board can be integrated. In a slot-controlled For example, the valve may be a slot in the cylinder wall act. In the field of lamella valves may also be a Construction act, in which the suction lamella sits in the piston and the suction takes place under the piston. In all of the above embodiments are constructively easy to implement versions of a inventive compressor.

The the inlet valve (s) associated end of a or each cylinder space can be a particular radially extending annular Extension, in particular the stroke of the suction lamella (s) limited and beveled to the attachment point of the suction plate (s) out or flattened.

In a further preferred embodiment of a compressor according to the invention, the ratio of piston diameter and piston stroke (D / s) in about 0.4 to 1.5, in particular 0.65 to 1.1, wherein a preferred value is about 0.95 , Additionally or alternatively, 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. Additionally or alternatively, 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. Again alternatively or additionally, the ratio of piston stroke to the stroke of the suction plate (s / t) 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 relates to the refrigerant R744 (CO ₂ ), it being noted at this point that for other refrigerants an adjustment of the parameter set is necessary and included in the spirit of the present invention.

In a compressor according to the invention also 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 α _max at low speeds of compressor. Preferably, the geometry and dimensioning of all translationally moving parts such as axial piston, piston rod or sliding blocks or the like. On the one hand and all rotationally moving parts such as swash plate, driver or the like. On the other hand such that for predetermined tilt angle of the swash plate, in particular between a predetermined minimum tilt angle and a predetermined maximum tilt angle, the moment M _{k, ges} due to the translationally moving masses, in particular the piston, optionally including sliding blocks, piston rods or the like. Is selected so smaller than the moment M _SW due to the Deviationsmoments, ie as the moment due to the inertia of the swash plate, that at high speeds of the compressor, in particular at very high speeds or at maximum speeds, the angle of maximum deflection of the swash plate is smaller than the angle α _max maximum deflection at lower speeds of the compressor. Also, such a structural design allows a minimization of the control interventions, especially in speed jumps with a simultaneous cost-effective production.

The Invention will be described below with regard to further advantages and Features by way of example and with reference to the enclosed Drawings described. The drawings show in:

1 a schematic representation to illustrate the derivation or calculation of the moments due to the translational and rotationally moving masses;

2 the swashplate mechanism of a compressor according to the invention in Explosionsdar position;

3a the torque distribution in a compressor according to the invention as a function of the tilt angle of the swash plate or the swivel ring;

3b the differential pressure between the engine room and the suction gas side or the pressure in the engine room as a function of the geometric displacement for a specific operating point of the compressor according to the invention at different compressor speeds;

4a C schematic explanations of the operation of a compressor according to the invention;

5 a representation of the differential pressure between the engine room and the suction gas side as a function of the mass flow of a compressor according to the invention;

6 four indicator diagrams for two operating points of the compressor according to the invention;

7 a schematic representation illustrating the design of the intake valve and the intake valves and the compressor geometry; and

8th the behavior of a compressor according to the invention at a speed jump from 2000 rev / min to 6000 rev / min.

In the method for controlling the refrigerant mass flow of a compressor, a torque 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 α _{g1 of} a swivel disk, which is in the form of a swivel ring 1 (see. 2 ) is present. For a compressor using CO ₂ as a refrigerant, the following parameter set is obtained, which approximately provides a moment equilibrium: m _k = 45 to 80 g (mass of the pistons), r _i = 17 to 25 mm (inner diameter of the swivel ring 1 ), r _a = 35 to 45 mm (outer diameter of the swivel ring 1 ), h = 12 to 17 mm and R = 24 to 35, where the material from which the swivel ring 1 is made of steel, gray cast iron or a bronze alloy. The design of the swashplate in the form of the swivel ring 1 ensures that the above-described ratio of M _SW and M _{k, ges can be achieved} in a structurally simple manner, since the center of gravity is located in the tilting joint on the drive shaft axis and not in dependence of the tilt angle of the pivot ring 1 varied. As a result, the control characteristics have a relatively linear course. A clarification of the derivation of the two above-described moments is out 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. In the case of complex geometries, in particular of the swash plate (if the illustrative consideration no longer yields satisfactory values), 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_yz = –J₁cosα₂cosα₃ – J₂cosβ₂cosβ₃ – J₃cosγ₂cosγ₃
• α₁ = 0
• β₁ = 90° – Richtungswinkel der x-Achse
• γ₁ = 90° – gegenüber den Hauptträgheitsachsen
  

• α₂ = 90°
• β₂ = Ψ – Richtungswinkel der y-Achse
• γ₂ = 90° + Ψ – gegenüber den Hauptträgheitsachsen
  

• α₃ = 90°
• β₃ = 90° – Ψ – Richtungswinkel der z-Achse
• γ₃ = Ψ – gegenüber den Hauptträgheitsachsen
  

In the simplified, but illustrative derivation of the mass moment of inertia, it is assumed that the center of gravity of the swash plate in the tilting joint is approximately on the shaft center axis, so no Steiner share or the like must be considered. For the derivation of the moment of deviation, the following mathematical relationships generally apply, whereby the relevant coordinate system in 1 is shown:

• J = -J _{yz 1} cos ₂ cos ₃ - J ₂ cosβ ₂ cosβ ₃ - J ₃ cosγ ₂ cosγ ₃
• α ₁ = 0
• β ₁ = 90 ° - direction angle of the x-axis
• γ ₁ = 90 ° - with respect to the principal axes of inertia
  

• α ₂ = 90 °
• β ₂ = Ψ - directional angle of the y-axis
• γ ₂ = 90 ° + Ψ - with respect to the principal axes of inertia
  

• α ₃ = 90 °
• β ₃ = 90 ° - Ψ - Directional angle of the z-axis
• γ ₃ = Ψ - with respect to the principal axes of inertia
  

The coordinate system used here proceeds as mentioned above 1 out. Furthermore applies to a "ring":

(Note: J ₃ ≈ 2 J ₂ )

For the moment of deviation, which is decisive for the pivoting movement, applies J Y Z = -J 2 cosΨ sinΨ + J 3 cosΨ sinΨ

sowie für das Moment M_sw infolge des Deviationsmoments:

Independent of 1 applies to the moment due to mass forces of the pistons:

and for the moment M _sw due to the moment of deviation:

In connection with the invention, the following torque ratio should be realized constructively for any tilt angle or tilt angle range:
M _SW ≥ M _{k, ges} or preferably the sub-case M _SW = M _{k, ges} or preferably (for CO ₂ as refrigerant) M _SW ≈ M _{k, ges}

Thus also applies:

As already explained, 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
M _SW ≥ M _{k, ges} or else the sub-case M _SW = M _{k, ges} holds.

θ

Drehwinkel der Welle (wobei die vor- und nachstehenden Betrachtungen der Einfachheit halber für θ = 0 angestellt werden)

η

Anzahl der Kolben

R

Abstand der Kolbenachse zur Wellenachse

ω

Wellendrehzahl

α

Kippwinkel des Schwenkringes/Schwenkscheibe

m_k

Masse eines Kolbens inklusive Gleitsteine bzw. Gleitsteinpaar

m_k,ges

Masse aller Kolben inklusive Gleitsteine

m_sw

Masse des Schwenkringes

r_a

Außenradius des Schwenkringes

r_i

Innenradius des Schwenkringes

h

Höhe des Schwenkringes

g

Dichte des Schwenkringes

V

Volumen des Schwenkringes

β_i

Winkelposition des Kolbens i

z_i

Beschleunigung des Kolbens i

F_mi

Massenkraft des Kolbens i (inklusive Gleitsteine)

M(F_mi)

Moment infolge der Massenkraft des Kolbens i

M_k,ges

Moment infolge der Massenkraft aller Kolben

M_sw

Moment infolge des Aufstellmomentes des Schwenkringes/Schwenkscheibe infolge des Deviationsmoments (J_yz)

In connection with the given equations,

θ

Angle of rotation of the shaft (the above and following considerations are for simplicity θ = 0)

η

Number of pistons

R

Distance of the piston axis to the shaft axis

ω

Shaft speed

α

Tilt angle of the swivel ring / swashplate

m _k

Mass of a piston including sliding blocks or pair of sliding blocks

m _{k, sat}

Mass of all pistons including sliding blocks

_sw

Mass of the swivel ring

r _a

Outer radius of the swivel ring

_i

Inner radius of the swivel ring

H

Height of the swivel ring

G

Density of the swivel ring

V

Volume of the swivel ring

β _i

Angular position of the piston i

z _i

Acceleration of the piston i

F _mi

Mass force of the piston i (including sliding blocks)

M (F _mi )

Moment due to the mass force of the piston i

M _{k, sat}

Moment due to the mass force of all pistons

M _sW

Moment due to the installation torque of the swivel ring / swashplate as a result of the deviation moment (J _yz )

An example of an engine in which a torque equilibrium (M _{k, ges} ≈ M _SW ) for at least one deflection angle α _{g1 of} the swash plate or the swivel ring 1 is prepared, ie for a compressor in which the torque balance feature of the method according to the invention is implemented is in 2 shown. However, before going into detail, it should be noted that a preferred embodiment of a compressor according to the invention (not shown in the drawings) comprises a housing, a cylinder block and a cylinder head. In the cylinder block pistons are mounted axially movable back and forth. The drive of the compressor via a pulley by means of a drive shaft 2 , The compressor described here is a variable piston type compressor, the piston stroke being controlled by a pressure differential defined by the pressures p _{s *} and p _c . Depending on the size of the pressure difference (this is time-dependent or dependent on the angle of rotation of the drive shaft 2 ) is a swash plate in the form of a swivel ring 1 more or less deflected or pivoted out of its or its vertical position (see also 3b If the pressure difference is large, then the tilting or deflection angle of the swivel ring is 1 small, while at small pressure differences, 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.

Out 2 It can be seen that the swash plate mechanism of the preferred embodiment, the swivel ring 1 , the drive shaft 2 , a sliding sleeve 3 on the drive shaft 2 axially against the action of at least one elastic element in the form of an annular or helical fitting or return spring 4 (Which preferably has a spring rate of C = 30 to 60 N / mm when using CO ₂ as a refrigerant), and a support element 5 and a power transmission element 6 includes. Alternatively, 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 element 5 is both radially and (in a direction perpendicular to the drive shaft axis) perpendicular to the force transmission element displaceable 6 hinged, which means that the support element 5 is slidably mounted in a plane (and not just along an axis). The support element 5 is cylindrical bolt-shaped and has a groove 7 on, by means of which the support element 5 with the power transmission element 6 is in operative engagement. This is the support element 5 facing end or is the support element 5 facing end portion of the power transmission element 6 formed 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 region is connected to the groove 7 of the support element 5 engaged. The advantage of the construction of the power transmission element 6 and the support element 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 middle area, the strength of the power transmission element decreases 6 too, while at its the drive shaft 2 facing end is sleeve-shaped. With the help of the sleeve-shaped part 8th of the power transmission element 6 is the same at the drive shaft 2 stored or attached. For a non-twist connection of the drive shaft 2 with the sleeve-shaped part 8th of the power transmission element 6 provides a feather key 2a , It should be noted at this point that in the present preferred embodiment, the power transmission element 6 one-piece and one-piece with the sleeve-shaped part 8th is trained. Alternatively, it could of course be the power transmission element 6 and the sleeve-shaped part 8th to act on two different components (possibly even from different materials). It should also be noted at this point that the force transmission element 6 or the sleeve-shaped part 8th of the power transmission element 6 two recesses in the form of grooves 9 having. It should be noted at this point that the force transmission element 6 and the sleeve-shaped part 8th also in one piece with the drive shaft 2 can be executed. This may, for example, be a forged part; a one-piece design is preferred for mass production.

As the inner diameter of the spring 4 larger than the outer diameter of the sleeve-shaped part 8th of the power transmission element 6 , can the sleeve-shaped part 8th in the assembled state of the swashplate mechanism under the spring 4 be pushed. So that means that the sleeve-shaped part 8th over the drive shaft 2 slipped and radially through the spring 4 on the drive shaft 2 is fixed. On the pen 4 opposite side of the power transmission element 6 then becomes the sliding sleeve 3 , which is a force transmission element 6 corresponding recess 10 has, over the drive shaft 2 slipped (sliding seat). The sliding sleeve 3 also has two recesses in the form of holes 11 on. Axial become the power transmission element 6 as well as the sliding sleeve 3 by a groove nut (not shown) on the drive shaft 2 secured, with the sliding sleeve 3 on the drive shaft 2 can reciprocate in the axial direction. The sleeve-shaped part 8th the power transmission element is connected to the spring 4 on the drive shaft 2 fixed against rotation. For a better starting behavior of the compressor is on the drive shaft 2 also a plate spring 12 arranged, which ensures that the compressor is not at a minimum deflection angle of the pivot ring 1 starts. Further, on the drive shaft 2 Stops in the form of stop discs 13 . 14 arranged, which limit the deflection angle of the pivot ring. The stop disc 13 serves as a stop for a minimum deflection angle while the stop disc 14 as a stop for a maximum deflection angle of the swivel ring 2 serves. On the back can also be provided a bearing seat for the main thrust bearing.

The support element 5 is in a cylindrical recess in the form of a bore 15 in the swivel ring 1 stored. The hole 15 extends perpendicular to the drive shaft axis. The securing of the support element 5 in the swivel ring 1 done by means of two snap rings 16 ,

It should be noted at this point that the force transmission element 6 in the present preferred embodiment, rotationally fixed to the drive shaft 2 connected is. It should also be noted at this point that through the sleeve-shaped training or the sleeve-shaped part 8th of the power transmission element 6 the drive shaft 2 is not broken and thus has corresponding stability. The clear width of the bore of the swivel ring 1 is at least slightly larger than the corresponding extent of the force transmission element 6 (Mountability).

In the present preferred embodiment, the mechanism is of support member 5 and power transmission element 6 not intended to torque from the shaft to the swash plate in the form of the swivel ring 1 transferred to. The bearings between support element 5 and power transmission element 6 , between power transmission element 6 and drive shaft 2 and between support element 5 and swivel ring 1 are not designed to transmit torque. It therefore eliminates a kind of driving function for the support element 5 and the power transmission element 6 , This is deliberately chosen for reasons of hysteresis, ie the tilting of the swivel ring 1 and the torque transfer are functionally decoupled from each other. The mechanism of power transmission element 6 and support element 5 essentially absorbs the piston forces. The torque in turn is from the drive shaft 2 to the swivel ring 1 by a tilting joint provided on the drive shaft central axis (realized by drive bolts 15a ) transfer. The torque between the sliding sleeve 3 and the swivel ring 1 transmitting drive bolts 15a are on the swivel ring with snap rings 16a arrested or secured. The swivel ring 1 indicates flattening 17 which leads to flattening 18 on the sliding sleeve 3 corresponding. In principle, it is also conceivable in other embodiments that the sliding sleeve 3 eliminates and the torque transmission in any form between the drive shaft and swivel ring 1 takes place directly (eg 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.

By decoupling the torque transmission and the gas power support can be achieved that in addition to the possibility of the support element 5 and the power transmission element 6 correspondingly small to dimension, an optimized surface pressure, in particular between the power transmission element 6 and support element 5 as well as between support element 5 and swivel ring 1 be achieved. Thereby and by the invention-specific design of support element 5 and power transmission element 6 or by the invention-specific articulation between the power transmission element 6 and support element 5 a compact design of the compressor can be achieved.

In 3a is a qualitative representation of the preferred interpretation of the moments according to the equations used (see. 1 ), wherein in addition to the rotationally and translationally related moments (M _SW and M _{k, ges} ), the sum of the moments is shown. How one 3a can be found over a wide range of the tilt angle or deflection angle of the pivot ring 1 about a moment equilibrium, which in the present figure is a representation of the moments above the pivot angle α for any rotational speed n of the drive shaft 2 is.

In 3b is for a fixed spring rate of the return spring 4 and for fixed pressure ratios on the high pressure and suction sides of 130 and 35 bar for different speeds n the differential pressure prevailing between engine room and suction side, each considered for a compressor with moment equilibrium and with a constant suction gas density over the whole speed range ρ, 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 for each speed n of the drive shaft 2 be accepted constantly. In practice, with increasing speed n, in particular the suction pressure p _{s *} and thus the suction gas density ρ are lowered.

It should be noted at this point that in 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 3a represented, an engine with aufregelndem behavior or an engine designed with abregelndem behavior. The moments M _SW and M _{k, ges} or their relationship to each other would (n) only provide correspondingly, for CO ₂ as the refrigerant course M _SW ≈ M _{k, ges} or M _SW > M _{k, ges is} preferred.

Let us consider the case M _SW + M _{k, ges} ≈ 0. Out 3b you can see that the control curves intersect in exactly one point. This is the starting point for the sum of the moments, where M _SW = M _{k, sat} . This point can be selected arbitrarily depending on the design of a compressor. Based on the equation of the sum of the moments 1 you can easily understand an influence of the tilt angle. The effect results from the courses of the terms tan (α) and sin (2α). That is, the sum of the moments can be balanced in the design for exactly one tilt angle. If this happens, for example, for the maximum tilt angle of the swash plate, there are smaller deviations in the sum of the moments for other tilt angles. These deviations can be kept relatively small.

For the following tilt angles, the setting of the torque balance makes sense:
for α _min ≤ α ≤ α _max
for α = (α _max - α _min ) / 2
for α = α _max , and
for α ≥ α _max

However, the last two cases are clearly to be preferred, it being noted that in the diagrams of the 3a and 3b approximately the third case (intersection at 95% of the maximum delivery volume), ie, a moment equilibrium for α = α _max , realized. In addition to the opposite to changes in speed in approximately neutral behaving engine, the tilt angle of the swivel ring 2 changed substantially by the variation of the pressures p _{s *} , p _d and the engine room pressure p _c . In a constant operating point for a given p and _{s *} p _d the change occurs substantially only by the drive mechanism chamber pressure p _c. The latter is also the diagrams of the 3b refer to.

In order to avoid or reduce control interventions, which are caused by changes in speed, in the method described here, the product of the compressor speed n, the Sauggasdichte ρ 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). This happens at least regionally, in particular for compressor speeds n between 600 and 9000 rpm and preferably for compressor speeds between 2500 and 7000 rpm. As a result, the delivery volume, ie the mass flow even without regulation of the engine room pressure p _c in about kept constant. This is simply illustrated by the following relationships: The delivery volume per time V [cm ³ / s] = V _geo [cm ³ ] × n [1 / s], where V _{geo stands} for the geometric delivery volume and n for the compressor speed. For the geometric delivery volume V _{geo the} following relationship applies: V _geo = D ² π / 4 × s × ŋ, where ŋ represents the number of pistons. The mass flow per time m [g / s] = V [cm ³ / s] × ρ [g / cm ³ ], where ρ stands for the time to be averaged Sauggasdichte in the cylinder chamber. This implies that m is a function of n, ρ, and s, where ρ is a function of t and p. From this it can be seen that m = K × s × n × ρ, where K is a constant, so that when the product of s, n and ρ is kept constant, the delivery volume or mass flow is kept constant.

Accordingly, such an inventive embodiment of the method with a moment equilibrium of the translational and rotational masses on the one hand and a constant maintenance of the above-described parameters on the other side allows a control of a compressor without readjusting the engine room pressure p _c for a corresponding operating point. About the control valve on the compressor and the system-side control is set only a desired operating point, while a change in the speed of the drive shaft 2 is largely intercepted by the self-compensating behavior of a compressor according to the invention. Apparently, this means that when a change in the speed of an approximately proportional behavior with respect to the delivered refrigerant mass flow must be realized. This means that even with a significant increase in the speed of the delivered refrigerant mass flow is to be kept constant or limited. This is apparativ (see this 7 ) solved in that in addition to the moment equilibrium, for example, by an embodiment of the pivot ring 2 and the engine according to 2 , at least one suction gas inlet valve is arranged, which is configured such that the refrigerant mass flow entering the cylinder bores is speed-dependent (in particular on the Sauggasdichte) 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. This means, in other words, that 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. It should also be mentioned at this point that when M _SW ≈ M _{k, ges} of the angle of deflection of the pivot ring 1 remains constant at speed jumps. In other words, when the rotational speed of the compressor doubles from n1 to n2, the refrigerant mass flow is kept substantially constant by a deflection angle reduction (here halving). This is caused by the lowering of the suction gas density (ρ ₁ × n ₁ × s ₁ ≈ ρ ₂ × n ₂ × s ₂ , where n ₂ = 2n ₁ and s ₁ × ρ ₁ ≈ 2 × ρ ₂ × s ₂ ).

The throttling by the suction-side valves is shown schematically in FIG 4a shown, wherein the influence of throttling by a log-ph diagram in 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. In addition to this is the refrigerant mass flow over the speed of the drive 2 -respectively. Compressor shaft in 4c applied (qualitative representation). In addition to the theoretical increase in the mass flow, 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.

As well as out 5 can be seen, can be effected due to the process control of the invention or design of a compressor that for speeds n1, n2, n3 and n4 at a respective doubling between the respective nearest rotational speeds (eg, n ₂ = 2 n ₁ ) halving of _Geo V × ρ can be represented without the pressure p _c in the engine room having to be changed, so that m = V _geo × ρ × n remains constant at said doubling of n. 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. In the embodiment of the compressor according to the invention must also be considered that losses due to the Sauggasdrosselung occur, which must be considered in the design of the valve geometry. These losses increase with increasing speed. The inventive design is achieved that V _geo × ρ changes so that the mass flow of the refrigerant remains constant, without the pressure in the engine room would have to be changed.

In 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. In other words, if the speed changes from n2 to n3, this means that the mass flow of the refrigerant has no losses from m2t to m3t is lifted. Due to the losses, however, the mass flow of the refrigerant is increased from a mass flow m2r to a mass flow m3r. As a result, it is conditional that an ideal course of the control behavior can be ensured by the design of the suction valves only over wide speed ranges and not over the entire speed range.

The background for this situation lies in the fact that, for example, when doubling the speed of the tilt angle of the swash plate would have to be halved. If the moments M _SW and M _{k, ges} were designed so that M _{SW was} greater than M _{k, ges} (overcompensation), then the speed would take a non-linear influence (this influence would be subject to a quadratic law). Similarly, the throttling has no linear (but a quadratic) influence. It makes sense to avoid at least one of these mutually reinforcing effects. In any case, it has already been mentioned in comparison to the prior art that overcompensation by the parameter M _SW would have to be paid for by a correspondingly unattractive component mass and component size. Accordingly _{, Mk, ges} M _{SW is} to be preferred.

Therefore it is advantageous, the control behavior in a medium speed range vote. In the speed range below is then a slight Undercompensation before, in the upper speed range is a slight overcompensation in front.

Compensation means in this context that at a speed doubling the geometric displacement or the tilt angle or stroke of the piston is changed so automatically that the mass flow of the refrigerant does not change substantially, the engine room pressure p _{c must} not be changed by a control intervention and the plant-side pressure level on the suction side or the high-pressure side as well as the cooling capacity does not change substantially due to the unaltered mass flow. With a constant refrigerant mass flow m = const (where, as explained above, m = V _geo × ρ × n), a change in the rotational speed n is effected by the fact that m = K × s × n × ρ, directly and in a timely manner on ρ × s, where ρ also acts on s. Accordingly, a very fast control response to the change of the rotational speed n is obtained. This would not be possible by means of the controlled system of the control valve, so that an overshoot can be avoided. The swivel ring 1 is therefore a kind of internal controller (watt controller).

under compensation is called in this context, that at Doubling the speed of the geometric displacement or the tilt angle or the stroke of the piston is changed so automatically that the mass flow of the refrigerant across from the starting position is somewhat reduced. A corrective control action becomes necessary. Likewise, overcompensation means in this context, that at Doubling the speed of the geometric displacement or the tilting angle or the stroke of the piston is changed in such a way automatically, that the Mass flow of the refrigerant across from the starting position slightly increased becomes. A corrective control intervention, as in the case of undercompensation, necessary.

An undercompensation of M _{k, ges} by M _SW or M _{k, ges} is not appropriate, since then the design of the suction valves would also eliminate this negative effect with.

That means summarized:
M _SW > M _{k, ges} (overcompensation of the translationally moving masses) causes disadvantages in the "package" area and undesirably increases the influence of the throttling by the suction valves.

M _SW <M _{k, ges} (undercompensation of the translationally moving masses) causes disadvantages in that the effect works against the intended effect of Sauggasdrosselung.

M _SW ≈ M _{k, tot} is ideal, the geometry of the suction valves and the compression chamber are adapted to it (especially for the refrigerant CO ₂ ).

Likewise, the terms over-, under- or compensation were used in addition to the relationship of the moments for the mass flow of the refrigerant:
Undercompensation: When the rotational speed n is doubled, the refrigerant mass flow m = V _geo × n × ρ, ie V _geo × ρ, is changed so automatically that the mass flow of the refrigerant relative to the starting position is reduced. A corrective control intervention is necessary.

Overcompensation: When doubling the speed is the geometric displacement so automatically changed that the Mass flow of the refrigerant across from the starting position increased becomes. Again, a corrective control intervention is necessary.

In Connection with speed jumps has been so far from increases the speed spoken. It goes without saying that the behavior is the same Form reflects at speed drops, in particular it is for example at. halving the speed of the target, changes the mass flow of the refrigerant automatic to compensate. In addition increased the geometric delivery volume accordingly.

In 6 Indicator diagrams are shown for two operating points to the influence of the valve losses as a function of the speed n of the drive shaft 2 to show. 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 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} greater than M _{k, ges} selected (overcompensation of the moments) in order to achieve a compensation in the range of the mass flow of the refrigerant. It should be noted at this point that the concrete method described here for CO ₂ as a refrigerant is transferable to any refrigerant, this transfer is included in the spirit of the present invention. In the two diagrams above, starting with a lower oscillator (lower left) in the suction pressure at which the suction valve opens, you can look at a compressor cycle. The gas is compressed in the sequence (counterclockwise process direction) until the outlet valve opens after a slight overshoot. After the outlet of the gas follows a re-expansion, which eventually leads back to the starting point.

The compressor has (see. 7 ) at the inlet side for the suction gas into the cylinder space a valve plate 19 with a suction lamella attached underneath 20 on. The suction lamella 20 is tongue-shaped and serves to control the Sauggaseinlasses. If the gas in the cylinder is compressed, the suction lamella closes 20 a passage throttle bore 21 while the suction lamella 20 during suction of the suction gas (due to the prevailing in the cylinder vacuum) by a stroke t (indicated by arrows 22 ) moved down and the refrigerant to be sucked or the suction gas through the passage throttle bore 21 Granted inlet into the cylinder.

The passage throttle bore 21 has a diameter d. Due to the geometry of the intake 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 also the geometry of the compressor, over wide operating ranges of the compressor according to the invention, results in a desired lowering of the suction pressure p _s . 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 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 a bore (passage throttle 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. A preferred parameter set is: N = 6; D = 18; s = 19; V = 4.7; t = 1.1; Accordingly, the energetically favorable variables describing the geometry of the compressor are a ratio of piston diameter and piston stroke of approximately 0.65 to 1.1, a ratio of piston diameter and passage throttle bore 21 in the valve plate 19 from about 2.5 to 4, a ratio of port 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 24.

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 of 0.4 to 1.5 for the ratio of piston diameter and piston stroke and values of 1 , 5 to 5 for the ratio of piston diameter and Durchgangsdrosselbohrung and values of 2.5 to 8 for the ratio of Durchgangsdrosselbohrung and stroke of the suction plate and values of about 10 to 30 for the ratio of piston stroke to lift the suction plate are energetically favorable. In this preferred embodiment, the passage throttle bore 21 used on the suction side as a throttle point and designed specifically in connection 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 suction valve configuration, the pressure p _{s *} sets, which ensures an optimal control behavior of the compressor.

In 8th 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 this behavior prevails.

Of the ideal area for the design is, as already mentioned, the average speed range, so for the above mentioned sizes short-term changes (Mass flow of the refrigerant as well as the pressures in the engine room at the suction side of the compressor and the pressure side of the compressor) are compensated for self-regulating.

This is also easy to test. With a compressor for the application of the refrigerant R744, 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} is understand by measuring and weighing. A measuring 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. The pressure p _s can be indexed in an indicator diagram and easily calculated by its relationship with the suction gas density ρ (ρ = f (p, T)). Accordingly, all sizes are determinable.

By the self-regulation according to the invention can be used in contrast to the prior art, a simple switching valve which is the gas flow from the high pressure side into the engine room can influence. The switching valve can then intervene when a another operating point should be set. An intervention on the control valve by a so-called feedback as in the state of Technology is not necessary. The control valve, which controls the gas flow from the pressure side of the compressor into the engine room of the compressor rules, must therefore no additional Signal supplied as is known in the art. At the booth The technique serves as additional Signals e.g. the change the mass flow of the refrigerant, the change a pressure difference, the change of suction pressure etc. For a once set operating point can self-regulation Fluctuations in the refrigerant mass flow due to the speed compensate. It is at this point on it pointed out that it it is essential that not only the mass flow can be kept substantially constant, but at the same time the pressure 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 set shall be. A so-called switching valve is opposite the stand The technique thus characterized in that the feedback area omitted can. Such a switching valve is much cheaper than the state of the art the technique used valves. Such a simple in a compressor according to the invention used valve is preferably a valve of the type, as they for today ABS or ESP valves are used.

It It should also be noted that the control according to the invention significantly faster than the previous scheme works. The ones to be regulated Become sizes about coincided with the increase in speed, according to the In the prior art, this happens with a time delay, since first a feedback variable can be tapped off must, that supplied to the control valve or assigned.

Even though the invention based on embodiments is described with a fixed combination of features, but includes also the conceivable further advantageous combinations of these features, as specifically, but not exhaustively, indicated by the subclaims are. All Features disclosed in the application documents are considered to be essential to the invention as far as they are individually or in combination with respect to State of the art are new.

1

swivel

2

drive shaft

2a

adjusting spring

3

sliding sleeve

4

Return spring

5

support element

6

Power transmission element

7

groove

8th

sleeve-shaped part of the power transmission element 6

9

groove

10

recess

11

drilling

12

Belleville spring

13 14

stop disc

15

drilling

15a

drive pin

16 16a

snap ring

17 18

flattening

19

valve plate

20

suction plate

21

Passage orifice

22

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Claims (20)

Method for regulating the refrigerant mass flow of a compressor, in particular axial piston compressor, furthermore in particular for motor vehicle air conditioning systems, furthermore in particular for CO ₂ as refrigerant, which has a drive shaft ( 2 ) and in its inclination to the drive shaft adjustable swash plate ( 1 ) which is arranged in a substantially defined by a housing of the compressor engine room and which by their deflection angle with respect to the drive shaft ( 2 ) Defines the piston stroke of the compressor, wherein approximately a moment equilibrium between a conditional by rotationally moving masses moment M _SW and a conditional by translationally moving masses moment M _{k, ges} for at least one deflection angle α _{g1 of} the swash plate ( 1 ) and wherein the product of the compressor speed n, the Sauggasdichte ρ and the piston stroke s for different compressor speeds n at least partially, in particular for compressor speeds n between 600 and 9000 U / min, in particular for compressor speeds between 2500 and 7000 U / min, automatically in is kept approximately constant, while a ruling in the engine room pressure p _{c is} also kept approximately constant (ρ × s × n = constant). Compressor according to claim 1, characterized in that in addition to the product of the compressor speed n, the Sauggasdichte ρ and the piston stroke s and the pressure p _d on a high pressure side of the compressor and / or the pressure p _s on a suction gas side of the compressor and / or a Regelventilstellgröße a control valve between the high pressure side and the engine room are kept approximately constant. Method according to Claim 1 or 2, characterized in that the moment equilibrium between M _SW and M _{k, ges is produced} at least for α _g1 = (α _max -α _min ) / 2, where α _{max is} the maximum deflection angle of the swash plate ( 1 ) and α _min the minimum deflection angle of the swash plate ( 1 ) designated. Method according to one of the preceding claims, characterized in that the instantaneous balance between M _SW and M _{k, ges is produced} at least for a deflection angle α _g1 , for which the following applies: α _min ≤ α _g1 ≤ α _max . Method according to one of the preceding claims, characterized in that the moment equilibrium between M _SW and M _{k, ges is established} at least for a deflection angle α _g1 = α _max . Method according to one of the preceding claims, characterized in that the moment equilibrium between M _SW and M _{k, ges is established} at least for a (notional) deflection angle α _g1 ≥ α _max . Method according to one of the preceding claims, characterized in that the pressure in the engine room p _{c is} varied in order to obtain a desired operating point, ie a desired refrigerant mass flow in the compressor. Method according to one of the preceding claims, characterized in that, with an increase in the rotational speed n, the suction pressure p _{s *} and thus ρ is lowered in such a way that the product of n, ρ and s is approximately constant. Compressor, in particular axial piston compressor, further in particular compressor for the air conditioning of a motor vehicle, with a housing and a substantially disposed in the housing, via a drive shaft ( 2 ) driven compressor unit for sucking and compressing a refrigerant with a likewise arranged in the housing swash plate ( 1 ), characterized in that the moments M _{SW are} due to the rotationally moving masses and M _{k, ges} due to the translationally moving masses in a predetermined ratio to each other, that the compression comprises at least one suction gas side arranged inlet valve which is configured so that it in the cylinder space reaching refrigerant mass flow speed-dependent influenced so that the ratio of the moments M _SW and M _{k, ges} on the one hand and the throttle performance of the at least one inlet valve on the other hand in relation to each other in such a way that at least over parts of the speed range, in particular for compressor speeds n between 600 and 9000 U / min, in particular for compressor speeds between 2500 and 7000 U / min of the pumped into the system refrigerant mass flow is approximately constant. Compressor according to Claim 9, characterized in that the at least one inlet valve is a pressure-controlled flap valve ( 19 . 20 . 21 ). Compressor according to claim 9 or 10, characterized that this at least one inlet valve slit-controlled valve is. Compressor according to one of claims 9 to 11, characterized in that the inlet valve, in particular lamellar valve, a valve plate ( 19 ) with passage throttle bore ( 21 ) and in particular a tongue-shaped suction lamella ( 20 ). Compressor according to one of claims 9 to 12, comprising a cylinder block and at least one, in particular 5 to 9 pistons, which are axially back and forth in provided in the cylinder block bores, characterized in that each cylinder is associated with an inlet valve and / or the corresponding Suction lamellae ( 20 ) are integrated in a Sauglamellenplatine. Compressor according to Claims 9 to 13, characterized in that the end of the or each cylinder chamber associated with the inlet valve comprises a radially extending annular enlargement which in particular covers the stroke of the suction lamella (s) ( 20 ) and to the attachment point of the suction lamella (s) ( 20 ) Beveled or flattened. Compressor according to one of claims 9 to 14, if after a the claims 8 to 10 additional with a cylinder block and at least one, in particular 5 to 9 pistons, which in axially provided in the cylinder block bores are reciprocally movable, characterized in that the ratio of Piston diameter and piston stroke (D / s) in about 0.4 to 1.5, in particular 0.65 to 1.1, further particularly in about 0.95. Compressor according to one of claims 9 to 15, in particular Claim 15, characterized in that the ratio of piston diameter and passage throttle bore in the valve plate (D / d) in about 1.5 to 5, in particular 2.5 to 4, in particular in about 3.6 is. Compressor according to one of claims 9 to 16, in particular Claim 15 or 16, characterized in that the ratio of passage throttle bore in the valve plate and stroke of the suction plate (d / t) in about 2.5 to 8, in particular 3.7 to 6.7, in particular in about 4.55 is. Compressor according to one of claims 9 to 17, in particular Claim 15 or 16 or 17, characterized in that the ratio of Piston stroke for lifting the suction plate (s / t) in about 10 to 30, in particular 14 to 24, further particularly in about 17.3. Compressor according to one of claims 9 to 18, characterized in that the tilting behavior of the swivel disk ( 1 ) is so automatically limiting effect that at high speeds of the compressor, especially at very high speeds or the maximum speed, the angle of maximum deflection of the swash plate ( 1 ) is smaller than the angle of maximum deflection α _max at low speeds of the compressor. Compressor according to one of claims 9 to 19, characterized in that the geometry and dimensioning of all translationally moving parts, such as axial piston, piston rod or sliding blocks, od. Like., On the one hand and all rotationally moving parts, such as swash plate ( 1 ), Take away od. Like., On the other hand are such that for predetermined tilt angle of the swash plate ( 1 ), in particular between a predetermined minimum tilt angle and a predetermined maximum tilt angle, the torque M _{k, ges} due to the translationally moving masses, in particular the piston, if necessary including sliding blocks, piston rods od. Like., Is chosen so smaller than the moment M _SW due the Deviationsmoments, ie as the moment due to the inertia of the swash plate, that at high speeds of the compressor, especially at very high speeds or at a maximum speed, the angle of maximum deflection α _{max of} the swash plate ( 1 ) is smaller than the angle of maximum deflection at lower speeds of the compressor.