EP2021630B1 - Method for controlling the refrigerant mass flow of a compressor - Google Patents

Description

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 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. Independent claims 1 and 9 are in two-part form over the disclosure of this document.

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 engine behavior, which makes the mass flow of a refrigerant compressor optimally controllable, with an optimal refrigerant mass flow, however, 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 optionally intervenes regulating. 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 determined by the length of the controlled system, with the controlled system being represented as follows: On a change in the compressor speed The tilt angle of the swash plate changes, which results in a change in the refrigerant mass flow, which results in a change in the ratio of p _d 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 inclined or swash plate, whereby an overcompensation of the translationally moved masses (pistons 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.

This object is achieved by a method having the features of claim 1, wherein advantageous developments and details of the invention are described in the subclaims.

In particular, 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 caused 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 _gl of the swash plate is brought about and in which the product of the compressor speed n, the Sauggasdichte p (the suction gas, which in the cylinder with a 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 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 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 is established between M _sw and M _{k, ges} at least for α _gl = (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 is established between M _sw and M _{k, ges} (at least) for a deflection angle α _gl for which α _{min ≦} α _gl ≦ α _max . Additionally or alternatively, the torque balance between M _sw and M _{k, ges} (at least) for a deflection angle α _gl = α _max and / or for a (fictitious) deflection angle α _gl ≥ α _max are 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 p are lowered such that the product of n, p 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.

With regard to the device-technical aspect, the object is achieved by a compressor with the features of claim 9.

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 Inlet 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 coordination of refrigerant mass flow loss-making 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, the inlet valve, which is responsible for the suction gas density p entering the cylinder, is a pressure-controlled lamella valve. Alternatively or additionally, 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 lamellar (or slot-controlled) valve, preferably has a valve plate with through-hole (s) or through-flow bore (s) and a particular tongue-shaped suction lamella. 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).

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. In turn 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 is located. 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 to be 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.

Fig. 1

eine schematische Darstellung zur Verdeutlichung der Herleitung bzw. Berechnung der Momente infolge der translatorisch und rotatorisch bewegten Massen;

Fig. 2

den Schwenkscheibenmechanismus eines erfindungsgemäßen Verdichters in Explosionsdarstellung;

Fig. 3a

die Momentenverteilung in einem erfindungsgemäßen Verdichter als Funktion des Kippwinkels der Schwenkscheibe bzw. des Schwenkrings;

Fig. 3b

den Differenzdruck zwischen Triebwerksraum und Sauggasseite bzw. den Druck im Triebwerksraum als Funktion des geometrischen Fördervolumens für einen bestimmten Betriebspunkt des erfindungsgemäßen Verdichters bei verschiedenen Verdichterdrehzahlen;

Fig. 4 a-c

schematische Erläuterungen zur Funktionsweise eines erfindungsgemäßen Verdichters;

Fig. 5

eine Darstellung des Differenzdrucks zwischen Triebwerksraum und Sauggasseite als Funktion des Massenstroms eines erfindungsgemäßen Verdichters;

Fig. 6

vier Indikatordiagramme für zwei Betriebspunkte des erfindungsgemäßen Verdichters;

Fig. 7

eine schematische Darstellung, welche die Auslegung des Einlaßventils bzw. der Einlaßventile und der Verdichtergeometrie verdeutlicht; und

Fig. 8

das Verhalten eines erfindungsgemäßen Verdichters bei einem Drehzahlsprung von 2000 U/min auf 6000 U/min.

The invention will now be described by way of example and with reference to the accompanying drawings in view of further advantages and features. The drawings show in:

Fig. 1

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

Fig. 2

the swashplate mechanism of a compressor according to the invention in an exploded view;

Fig. 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;

Fig. 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;

Fig. 4 ac

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

Fig. 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;

Fig. 6

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

Fig. 7

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

Fig. 8

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 regulating the refrigerant mass flow of a compressor, 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 α _{gl of} a swivel disk which is in the form of a swivel ring 1 (see. Fig. 2 ) is present. For a compressor using CO ₂ as the refrigerant, the following parameter set results, which provides approximately a moment equilibrium: m _k = 45 to 80 g (mass of the pistons), r _i = 17 to 25 mm (Inner diameter of the pivot ring 1), r _a = 35 to 45 mm (outer diameter of the pivot ring 1), h = 12 to 17 mm and R = 24 to 35, wherein the material from which the pivot ring 1 is made, steel, gray cast iron or a bronze alloy is. The design of the swash plate 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 mass in the tilting joint lies on the drive shaft axis and not in dependence of the tilt angle of the swivel ring 1 varies. As a result, the control characteristics have a relatively linear course. 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. 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.

α₁ = 0 β₁ = 90° Richtungswinkel der x-Achse γ₁ = 90° gegenüber den Hauptträgheitsachsen ξ, η, ·ζ α₂ = 90° β₂ = ψ Richtungswinkel der y-Achse γ₂ = 90° + ψ gegenüber den Hauptträgheitsachsen ξ, η, ·ζ α₃ = 90° β₃ = 90°-ψ Richtungswinkel der z-Achse γ₃ = ψ gegenüber den Hauptträgheitsachsen ξ, η, ·ζ 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 Fig. 1 is shown: $${\mathrm{J}}_{\mathrm{Y\; Z}}\mathrm{=}\mathrm{-}{\mathrm{<figure-callout\; id="1"\; label="J"\; filenames="imgf0002.png"\; state="\{\{state\}\}">J</figure-callout>}}_{\mathrm{1}}{{\mathrm{<figure-callout\; id="2"\; label="cos"\; filenames="imgf0002.png"\; state="\{\{state\}\}">cos</figure-callout>}}_2\cos }_3-{\mathrm{<figure-callout\; id="2"\; label="J"\; filenames="imgf0002.png"\; state="\{\{state\}\}">J</figure-callout>}}_2{{\mathrm{<figure-callout\; id="2"\; label="cos\beta "\; filenames="imgf0002.png"\; state="\{\{state\}\}">cos\beta </figure-callout>}}_2\mathrm{cos\beta }}_3-{\mathrm{<figure-callout\; id="3"\; label="J"\; filenames="imgf0002.png,imgf0006.png"\; state="\{\{state\}\}">J</figure-callout>}}_3{{\mathrm{<figure-callout\; id="2"\; label="cos\gamma "\; filenames="imgf0002.png"\; state="\{\{state\}\}">cos\gamma </figure-callout>}}_2\mathrm{<figure-callout\; id="3"\; label="cos\gamma "\; filenames="imgf0002.png,imgf0006.png"\; state="\{\{state\}\}">cos\gamma </figure-callout>}}_3$$

α ₁ = 0 β ₁ = 90 ° Directional angle of the x-axis γ ₁ = 90 ° opposite the main axes of inertia ξ, η, · ζ α ₂ = 90 ° β ₂ = ψ Direction angle of the y-axis γ ₂ = 90 ° + ψ opposite the main axes of inertia ξ, η, · ζ α ₃ = 90 ° β ₃ = 90 ° -ψ Direction angle of the z-axis γ ₃ = ψ opposite the main axes of inertia ξ, η, · ζ

sowie $${\mathrm{J}}_{\mathrm{3}}\mathrm{=}{\mathrm{J}}_{\mathrm{\zeta }}\mathrm{=}\frac{m}{\mathrm{2}}\left({{\mathrm{r}}_{\mathrm{a}}}^{\mathrm{2}}\mathrm{+}{{\mathrm{r}}_{\mathrm{i}}}^{\mathrm{2}}\right)$$

The coordinate system used here proceeds as mentioned above Fig. 1 out. Furthermore applies to a "ring": $${\mathrm{<figure-callout\; id="2"\; label="J"\; filenames="imgf0002.png"\; state="\{\{state\}\}">J</figure-callout>}}_2\mathrm{=}{\mathrm{J}}_{\mathrm{\eta }}\mathrm{=}\frac{\mathrm{<figure-callout\; id="4"\; label="m"\; filenames="imgf0002.png"\; state="\{\{state\}\}">m</figure-callout>}}{4}\left({{\mathrm{r}}_{\mathrm{a}}}^{\mathrm{2}}\mathrm{+}{{\mathrm{<figure-callout\; id="2"\; label="r\; i"\; filenames="imgf0002.png"\; state="\{\{state\}\}">r\; </figure-callout>}}_{\mathrm{i}}}^{\mathrm{2}}+\frac{{\mathrm{<figure-callout\; id="2"\; label="H"\; filenames="imgf0002.png"\; state="\{\{state\}\}">H</figure-callout>}}^2}{3}\right)$$

such as $${\mathrm{<figure-callout\; id="3"\; label="J"\; filenames="imgf0002.png,imgf0006.png"\; state="\{\{state\}\}">J</figure-callout>}}_{\mathrm{3}}\mathrm{=}{\mathrm{J}}_{\mathrm{\zeta }}\mathrm{=}\frac{\mathrm{<figure-callout\; id="2"\; label="m"\; filenames="imgf0002.png"\; state="\{\{state\}\}">m</figure-callout>}}{\mathrm{2}}\left({{\mathrm{r}}_{\mathrm{a}}}^{\mathrm{2}}\mathrm{+}{{\mathrm{<figure-callout\; id="2"\; label="r\; i"\; filenames="imgf0002.png"\; state="\{\{state\}\}">r\; </figure-callout>}}_{\mathrm{i}}}^{\mathrm{2}}\right)$$

(Note: J ₃ ≈ 2 J ₂ )

For the moment of deviation, which is decisive for the pivoting movement, the following applies: $${\mathrm{J}}_{\mathrm{Y\; Z}}\mathrm{=}\mathrm{-}{\mathrm{<figure-callout\; id="2"\; label="J"\; filenames="imgf0002.png"\; state="\{\{state\}\}">J</figure-callout>}}_{\mathrm{2}}\mathrm{cos\psi \; sin\psi }\mathrm{+}{\mathrm{<figure-callout\; id="3"\; label="J"\; filenames="imgf0002.png,imgf0006.png"\; state="\{\{state\}\}">J</figure-callout>}}_{\mathrm{3}}\mathrm{cos\psi \; sin\psi }$$

Z_i = Rω² tanα cosβ_i
F_mi = m_k z_i
M(F_mi) = m_k R cosβ_i z_i $${\mathrm{M}}_{\mathrm{k}\mathrm{,}\mathrm{ges}}\mathrm{=}{\mathrm{m}}_{\mathrm{k}}\mathrm{R}{\displaystyle \mathrm{\sum }_{\mathrm{i}\mathrm{=}\mathrm{1}}^{\mathrm{n}}}{\mathrm{z}}_{\mathrm{i}}\cos {\mathrm{\beta }}_{\mathrm{i}}$$

sowie für das Moment M_sw infolge des Deviationsmoments: $${\mathrm{M}}_{\mathrm{sw}}\mathrm{=}{\mathrm{J}}_{\mathrm{yz}}{\mathrm{\omega }}^{\mathrm{2}}$$

$${\mathrm{J}}_{\mathrm{yz}}\mathrm{=}\left\{\frac{\mathit{msw}}{2}\left({{\mathrm{r}}_{\mathrm{a}}}^{\mathrm{2}}\mathrm{+}{{\mathrm{r}}_{\mathrm{i}}}^{\mathrm{2}}\right)-\frac{\mathit{msw}}{4}\left({{\mathrm{r}}_{\mathrm{a}}}^{\mathrm{2}}\mathrm{+}{{\mathrm{r}}_{\mathrm{i}}}^{\mathrm{2}}+\frac{h^2}{3}\right)\right\}\mathrm{cos\alpha \; sin\alpha }$$

$${\mathrm{J}}_{\mathrm{yz}}\mathrm{=}\frac{\mathit{msw}}{\mathrm{24}}\sin \mathrm{2}\mathrm{\alpha }\left(\mathrm{3}{{\mathrm{r}}_{\mathrm{a}}}^{\mathrm{2}}\mathrm{+}\mathrm{3}{{\mathrm{r}}_{\mathrm{i}}}^{\mathrm{2}}\mathrm{-}{\mathrm{h}}^{\mathrm{2}}\right)$$

Independent of Fig. 1 applies to the moment due to mass forces of the pistons: $${\mathrm{\beta }}_{\mathrm{i}}\mathrm{=}\mathrm{\theta }\mathrm{+}\mathrm{2}\mathrm{\pi }\left(\mathrm{i}\mathrm{-}\mathrm{1}\right)\frac{1}{n}$$

Z _i = Rω ² tanα cosβ _i
F _mi = m _k z _i
M (F _mi ) = m _k R cosβ _i z _i $${\mathrm{M}}_{\mathrm{k}\mathrm{.}\mathrm{ges}}\mathrm{=}{\mathrm{m}}_{\mathrm{k}}\mathrm{R}{\displaystyle \underset{\mathrm{i}\mathrm{=}\mathrm{1}}{\overset{\mathrm{n}}{\mathrm{\Sigma }}}}{\mathrm{z}}_{\mathrm{i}}\cos {\mathrm{\beta }}_{\mathrm{i}}$$

and for the moment M _sw due to the moment of deviation: $${\mathrm{M}}_{\mathrm{sw}}\mathrm{=}{\mathrm{J}}_{\mathrm{Y\; Z}}{\mathrm{\omega }}^{\mathrm{2}}$$

$${\mathrm{J}}_{\mathrm{Y\; Z}}\mathrm{=}\left\{\frac{\mathit{<figure-callout\; id="2"\; label="msw"\; filenames="imgf0002.png"\; state="\{\{state\}\}">msw</figure-callout>}}{2}\left({{\mathrm{r}}_{\mathrm{a}}}^{\mathrm{2}}\mathrm{+}{{\mathrm{r}}_{\mathrm{i}}}^{\mathrm{2}}\right)-\frac{\mathit{<figure-callout\; id="4"\; label="msw"\; filenames="imgf0002.png"\; state="\{\{state\}\}">msw</figure-callout>}}{4}\left({{\mathrm{r}}_{\mathrm{a}}}^{\mathrm{2}}\mathrm{+}{{\mathrm{<figure-callout\; id="2"\; label="r\; i"\; filenames="imgf0002.png"\; state="\{\{state\}\}">r\; </figure-callout>}}_{\mathrm{i}}}^{\mathrm{2}}+\frac{{\mathrm{<figure-callout\; id="2"\; label="H"\; filenames="imgf0002.png"\; state="\{\{state\}\}">H</figure-callout>}}^2}{3}\right)\right\}\mathrm{cos\alpha \; sin\alpha }$$

$${\mathrm{J}}_{\mathrm{Y\; Z}}\mathrm{=}\frac{\mathit{msw}}{\mathrm{24}}\mathrm{<figure-callout\; id="2"\; label="sin"\; filenames="imgf0002.png"\; state="\{\{state\}\}">sin</figure-callout>}\mathrm{2}\mathrm{\alpha }\left(\mathrm{3}{{\mathrm{r}}_{\mathrm{a}}}^{\mathrm{2}}\mathrm{+}\mathrm{3}{{\mathrm{r}}_{\mathrm{i}}}^{\mathrm{2}}\mathrm{-}{\mathrm{<figure-callout\; id="2"\; label="H"\; filenames="imgf0002.png"\; state="\{\{state\}\}">H</figure-callout>}}^{\mathrm{2}}\right)$$

• M_SW ≥ M_k,ges bzw. bevorzugt der Unterfall M_SW = M_k,ges bzw. bevorzugt (für CO₂ als Kältemittel) M_SW ≈ M_k,ges

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}

bzw. für CO₂: $$\left[{\mathrm{\omega }}^{\mathrm{2}}{\mathrm{R}}^{\mathrm{2}}{\mathrm{m}}_{\mathrm{k}}\mathrm{tan\alpha }{\displaystyle \mathrm{\sum }_{i\mathrm{=}\mathrm{1}}^n}{\cos }^{\mathrm{2}}\mathrm{\beta }\approx {\mathrm{\omega }}^{\mathrm{2}}\frac{\mathit{msw}}{\mathrm{24}}\sin \mathrm{2}\mathrm{\alpha }\left({\mathrm{3}{\mathrm{r}}_{\mathrm{a}}}^{\mathrm{2}}\mathrm{+}{\mathrm{3}{\mathrm{r}}_{\mathrm{i}}}^{\mathrm{2}}\mathrm{-}{\mathrm{h}}^{\mathrm{2}}\right)\right]$$

Thus also applies: $$\left[{\mathrm{<figure-callout\; id="2"\; label="\omega "\; filenames="imgf0002.png"\; state="\{\{state\}\}">\omega </figure-callout>}}^{\mathrm{2}}{\mathrm{<figure-callout\; id="2"\; label="R"\; filenames="imgf0002.png"\; state="\{\{state\}\}">R</figure-callout>}}^{\mathrm{2}}{\mathrm{m}}_{\mathrm{k}}\mathrm{tan\alpha }{\displaystyle \underset{i\mathrm{=}\mathrm{1}}{\overset{\mathrm{<figure-callout\; id="2"\; label="n\; cos"\; filenames="imgf0002.png"\; state="\{\{state\}\}">n\; </figure-callout>}}{\mathrm{\Sigma }}}}{\cos }^{}{}^{\mathrm{2}}\mathrm{\beta }\mathrm{\le }{\mathrm{\omega }}^{\mathrm{2}}\frac{\mathit{msw}}{\mathrm{24}}\mathrm{<figure-callout\; id="2"\; label="sin"\; filenames="imgf0002.png"\; state="\{\{state\}\}">sin</figure-callout>}\mathrm{2}\mathrm{\alpha }\left({\mathrm{3}{\mathrm{r}}_{\mathrm{a}}}^{\mathrm{2}}\mathrm{+}{\mathrm{3}{\mathrm{r}}_{\mathrm{i}}}^{\mathrm{2}}\mathrm{-}{\mathrm{<figure-callout\; id="2"\; label="H"\; filenames="imgf0002.png"\; state="\{\{state\}\}">H</figure-callout>}}^{\mathrm{2}}\right)\right]$$

or for CO ₂ : $$\left[{\mathrm{<figure-callout\; id="2"\; label="\omega "\; filenames="imgf0002.png"\; state="\{\{state\}\}">\omega </figure-callout>}}^{\mathrm{2}}{\mathrm{<figure-callout\; id="2"\; label="R"\; filenames="imgf0002.png"\; state="\{\{state\}\}">R</figure-callout>}}^{\mathrm{2}}{\mathrm{m}}_{\mathrm{k}}\mathrm{tan\alpha }{\displaystyle \underset{i\mathrm{=}\mathrm{1}}{\overset{\mathrm{<figure-callout\; id="2"\; label="n\; cos"\; filenames="imgf0002.png"\; state="\{\{state\}\}">n\; </figure-callout>}}{\mathrm{\Sigma }}}}{\cos }^{}{}^{\mathrm{2}}\mathrm{\beta }\approx {\mathrm{\omega }}^{\mathrm{2}}\frac{\mathit{msw}}{\mathrm{24}}\mathrm{<figure-callout\; id="2"\; label="sin"\; filenames="imgf0002.png"\; state="\{\{state\}\}">sin</figure-callout>}\mathrm{2}\mathrm{\alpha }\left({\mathrm{3}{\mathrm{r}}_{\mathrm{a}}}^{\mathrm{2}}\mathrm{+}{\mathrm{3}{\mathrm{r}}_{\mathrm{i}}}^{\mathrm{2}}\mathrm{-}{\mathrm{<figure-callout\; id="2"\; label="H"\; filenames="imgf0002.png"\; state="\{\{state\}\}">H</figure-callout>}}^{\mathrm{2}}\right)\right]$$

As already explained, the (tilt) moment of the swashplate can be deliberately adjusted by various parameters (geometry, density distribution, mass, center of mass) as a result of the associated deviation moment

M _sw ≥ M _{k, ges} or else the subcorruption 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 _{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. 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 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 , Depending on the size of the pressure difference (this is time-dependent or dependent on the rotation angle of the drive shaft 2) is 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.

Out Fig. 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 a helical or return spring 4 (the in one application of CO ₂ as refrigerant preferably has a spring rate of C = 30 to 60 N / mm has), and a support member 5 and a power transmission element 6 comprises. 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 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. For this purpose, 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. It should be noted at this point that in the present preferred embodiment, the power transmission element 6 is integrally formed and also einstoffig with the sleeve-shaped part 8. Alternatively, it could, of course, the power transmission element 6 and the sleeve-shaped part 8 by two different components (possibly even of different materials) act. It should also be noted at this point that 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. It should be noted at this point that the power transmission element 6 and the sleeve-shaped part 8 can also be designed in one piece with the drive shaft 2. This may, for example, be a forged part; a one-piece design is preferred for mass production.

Since the inner diameter of the spring 4 is greater than the outer diameter of the sleeve-shaped part 8 of the force transmission element 6, the sleeve-shaped part 8 in assembled state of the swashplate mechanism are pushed 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. On the side facing away from the spring 4 of the force transmission element 6, 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. Axially, 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. For a better starting behavior of the compressor, 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. Further, 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, while 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 backup of the support member 5 in the pivot ring 1 by means of two snap rings 16th

It should be noted at this point that 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).

In the present preferred embodiment, 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 support member 5 and power transmission element 6, between the power transmission element 6 and drive shaft 2 and between support member 5 and pivot ring 1 are not designed to transmit torque. Accordingly, it eliminates a kind of driving function for the support member 5 and the power transmission element 6. This is deliberately chosen for reasons of hysteresis, ie the tilting of the pivot 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. In principle, it is also conceivable in other embodiments that the sliding sleeve 3 is dispensed with and the torque transmission takes place directly in any desired form between the drive shaft and the swivel ring 1 (eg via flattenings 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 ability to dimension the support member 5 and the power transmission element 6 correspondingly small, an optimized surface pressure, in particular between the power transmission element 6 and support member 5 and between support member 5 and pivot ring 1 can be achieved. Thereby, and by the invention-specific design of support member 5 and power transmission element 6 and by the invention-specific articulation between the power transmission element 6 and support member 5, a compact design of the compressor can be achieved.

In 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, which is in the present figure is a representation of the moments on the pivot angle α for any speed n of the drive shaft 2.

In 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.

It should be noted at this point that in 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 aufregelndem 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.

In the following, the case M _sw + M _{k, ges} ≈ 0 is considered. Out Fig. 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 Fig. 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.

• für α_min ≤ α ≤ α_max
• für α = (α_max - α_min)/2
• für α = α_max, und
• für α ≥ α_max

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 two latter cases are clearly to be preferred, it being noted that approximately the third case (intersection at 95% of the maximum delivery volume), ie a moment equilibrium for α = α _max , is realized in the diagrams of FIGS. 3 a and 3 b , In addition to the relative to changes in rotational speed in approximately neutral behaving engine, the tilt angle of the swivel ring 2 is essentially changed by the variation of the pressures p _s *, p _d and the engine room pressure p _c . At a constant operating point at given p _s * and p _d , the change occurs essentially only by the engine room pressure p _c . The latter is also the diagrams of the Fig. 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 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). 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, can be kept approximately constant even without regulation of the engine room pressure p _c . This is simply illustrated by the following relationships: 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. For the geometric delivery volume V _{geo the} following relationship applies: V _geo = D ² π / 4 xsx η, where η represents the number of pistons. The mass flow per time m [g / s] = V [cm ³ / s] x ρ [g / cm ³ ] where p stands for the time to be averaged Sauggasdichte in the cylinder chamber. It follows that m is a function of n, p and s, where p is a function of t and p. It can be seen that m = K xsxnx ρ, where K is a constant, so that when the product of s, n and p is kept constant, the delivery volume or the mass flow are 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. Only a desired operating point is set via the control valve on the compressor and the system-side control, while a change in the rotational speed of the drive shaft 2 is largely absorbed 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 Fig. 7 ) solved in that in addition to the instantaneous equilibrium, for example, by an embodiment of the pivot ring 2 and the engine according to Fig. 2 , 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 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 be further mentioned at this point that at M _SW ≈ M _{k, ges} the deflection angle of the pivot ring 1 remains constant in 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 xn ₁ xs ₁ ≈ ρ ₂ xn ₂ xs ₂ , where n ₂ = 2n ₁ and s, x ρ1 ≈ 2 x ρ ₂ xs ₂ ).

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. In addition to this is still the refrigerant mass flow over the speed of the drive 2 -bzw. Compressor shaft in Fig. 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 Fig. 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 ) a halving of V _geo x ρ can be represented without the pressure p _c in the engine room must be changed so that m = V _geo x ρ xn 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 xp changes so that the mass flow of the refrigerant remains constant, without the pressure in the engine room would have to be changed.

In 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. In other words, when the rotational speed is changed from n2 to n3, 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. 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 would be} greater than M _{k, saturated} (overcompensation), then the speed would have a nonlinear 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 with the prior art that overcompensation by the parameter M _sw would have to be paid for by means of a correspondingly unattractive component mass and component size. Accordingly, M _{k, ges} ≈ M _{sw is} to be preferred.

Therefore, it is advantageous to tune the control behavior in a medium speed range. In the speed range below there is then a slight undercompensation, in the upper speed range is slightly overcompensation.

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 x ρ xn), a change in the rotational speed n due to the fact that m = K xsxnx ρ holds is immediate and timely to ρ xs, 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).

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. Likewise, 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.

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.

• M_sw > M_k,ges (Überkompensation der translatorisch bewegten Massen) bewirkt Nachteile im Bereich "package" und verstärkt den Einfluß der Drosselung durch die Saugventile in unerwünschtem Maße.
• M_sw < M_k,ges (Unterkompensation der translatorisch bewegten Massen) bewirkt Nachteile dadurch, daß der Effekt gegen den gewollten Effekt der Sauggasdrosselung arbeitet.
• M_sw ≈ M_k,ges ist ideal, die Geometrie im wesentlichen der Saugventile und des Verdichtungsraums sind darauf abgestimmt (insbesondere für das Kältemittel CO₂).

That means summarized:

• M _sw > M _{k, ges} (overcompensation of the translationally moving masses) causes disadvantages in the "package" area and increases the influence of the throttling by the suction valves to an undesirable degree.
• 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, is ideal, the geometry of essentially the suction valves and the compression chamber are adapted to it (especially for the refrigerant CO ₂ ).

• Unterkompensation: Bei einer Verdoppelung der Drehzahl n wird der Kältemittel-Massenstrom m = V_geo x n x ρ, d.h. also V_geo x ρ, derart selbsttätig verändert, daß der Massenstrom des Kältemittels gegenüber der Ausgangslage verringert wird. Ein korrigierender Regeleingriff ist notwendig.
• Überkompensation: Bei einer Verdoppelung der Drehzahl wird das geometrische Fördervolumen derart selbsttätig verändert, daß der Massenstrom des Kältemittels gegenüber der Ausgangslage erhöht wird. Auch hier ist ein korrigierender Regeleingriff notwendig.

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 doubling the speed n, the refrigerant mass flow m = V _geo xnx ρ, ie V _geo x ρ, so automatically changed so that the mass flow of the refrigerant is reduced compared to the starting position. A corrective control intervention is necessary.
• Overcompensation: When the speed is doubled, the geometric delivery volume is automatically changed in such a way that the mass flow of the refrigerant relative to the starting position is increased. Again, a corrective control intervention is necessary.

In connection with speed jumps has been spoken of increases in speed. It is obvious that the behavior is reflected in the same form at speed drops, in particular it is for example. halving the speed of the target to compensate for changes in the mass flow of the refrigerant automatically. For this purpose, the geometric delivery volume increases accordingly.

In 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. It should be noted at this point that the concrete method described here for CO ₂ as a refrigerant can be applied to any refrigerant, this transmission in the spirit of the present invention is included. 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. 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. When the gas in the cylinder is compressed, 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 . 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 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. 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 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

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 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.

In 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 selbstregelnd.

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} 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. The pressure p _s can be in are indexed to an indicator diagram and easily calculated by its relation to the suction gas density ρ (ρ = f (p, T)). Accordingly, all sizes are determinable.

By the self-regulation according to the invention, contrary to the prior art, 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. In the prior art, as 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. For a once set operating point, 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.

It should also be noted that the inventive scheme works much faster than the previous scheme. The variables to be controlled are regulated at about the same time as the speed increases, according to the state of the art, this happens with a time delay, since first a feedback variable must be tapped, which is fed or assigned to the control valve.

However, while the invention will be described in terms of embodiments having a fixed combination of features, it also encompasses the conceivable further advantageous combinations of these features, which are particularly, but not exhaustively, indicated by the subclaims. All disclosed in the application documents Features are claimed as essential to the invention, as far as they are new individually or in combination over the prior art.

LIST OF REFERENCE NUMBERS

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 force 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

arrow

Claims (20)

1. Method of controlling the coolant mass flow of a compressor, especially of an axial piston compressor, moreover especially for motor vehicle air-conditioning systems, moreover especially for CO₂ as coolant, which compressor has a drive shaft (2) and a tilt plate (1) which is variable in terms of its inclination to the drive shaft and which is arranged in a drive mechanism chamber of the compressor substantially defined by a housing of the latter and which, by means of its angle of deflection with respect to the drive shaft (2), defines the piston stroke of the compressor,
   wherein there is brought about, approximately, a moment equilibrium between a moment M_sw caused by masses moved in rotation and a moment M_k,ges caused by masses moved in translation for at least one angle of deflection α_gl of the tilt plate (1),
   characterised in that
   the product of the compressor speed of rotation n, the inlet gas density p and the piston stroke s for different compressor speeds of rotation n is automatically kept approximately constant at least over a range, especially for compressor speeds of rotation n between 600 and 9000 rpm, especially for compressor speeds of rotation between 2500 and 7000 rpm, whilst a pressure p_c prevailing in the drive mechanism chamber is likewise kept approximately constant (ρ x s x n = constant).
2. Method according to claim 1,
   characterised in that
   in addition to the product of the compressor speed of rotation n, the inlet gas density ρ and the piston stroke s, there are also kept approximately constant the pressure p_d at the high-pressure side of the compressor and/or the pressure p_s at the inlet gas side of the compressor and/or a control valve setting variable of a control valve between the high-pressure side and the drive mechanism chamber.
3. Method according to claim 1 or 2,
   characterised in that
   the moment equilibrium between M_sw and M_k,ges is produced at least for α_gl = (α_max - α_min)/2, wherein α_max denotes the maximum angle of deflection of the tilt plate (1) and α_min denotes the minimum angle of deflection of the tilt plate (1).
4. Method according to one of the preceding claims,
   characterised in that
   the moment equilibrium between M_sw and M_k,ges is produced at least for an angle of deflection α_gl such that: α_min ≤ α_gl ≤ α_max.
5. Method according to one of the preceding claims,
   characterised in that
   the moment equilibrium between M_sw and M_k,ges is produced at least for an angle of deflection α_gl = α_max.
6. Method according to one of the preceding claims,
   characterised in that
   the moment equilibrium between M_sw and M_k,ges is produced at least for a (hypothetical) angle of deflection α_gl ≥ α_max.
7. Method according to one of the preceding claims,
   characterised in that
   the pressure in the drive mechanism chamber p_c is varied in order to obtain a desired operating point, that is to say a desired coolant mass flow in the compressor.
8. Method according to one of the preceding claims,
   characterised in that,
   in the case of an increase in the speed of rotation n, the inlet pressure p_s* - and together therewith ρ - is lowered so that the product of n, ρ and s is approximately constant.
9. Compressor, especially an axial piston compressor, moreover especially a compressor for the air-conditioning system of a motor vehicle, having a housing and, for drawing in and compressing a coolant, a compressor unit arranged substantially in the housing and driven by means of a drive shaft (2) and having a tilt plate (1) likewise arranged in the housing,
   characterised in that
   the moments M_sw due to the masses moved in rotation and M_k,ges due to the masses moved in translation are in a predetermined ratio to one another,
   and in that the compressor has, arranged on the inlet gas side, at least one inlet valve which is so configured that it influences, in dependence on the speed of rotation, the coolant mass flow entering the cylinder space so that, on the one hand, the ratio of the moments M_sw and M_k,ges and, on the other hand, the throttling performance of the at least one inlet valve are in a relation to one another such that, at least over parts of the speed of rotation range, especially for compressor speeds of rotation n between 600 and 9000 rpm, especially for compressor speeds of rotation between 2500 and 7000 rpm, the coolant mass flow introduced into the system is approximately constant.
10. Compressor according to claim 9,
    characterised in that
    the at least one inlet valve is a pressure-controlled reed valve (19, 20, 21).
11. Compressor according to claim 9 or 10,
    characterised in that
    the at least one inlet valve is a slot-controlled valve.
12. Compressor according to one of claims 9 to 11,
    characterised in that
    the inlet valve, especially reed valve, has a valve plate (19) having a throttling through-bore (21) and an especially tongue-like inlet blade (20).
13. Compressor according to one of claims 9 to 12,
    having a cylinder block and at least one, especially 5 to 9, piston(s) which is/are arranged to move axially back and forth in bores provided in the cylinder block,
    characterised in that
    an inlet valve is associated with each cylinder, and/or the corresponding inlet blades (20) are integrated in an inlet blade plate.
14. Compressor according to claim 9 to 13,
    characterised in that
    that end of the or each cylinder space which is associated with the inlet valve has a radially extending annular extension, which limits especially the stroke of the inlet blade(s) (20) and which is bevelled off or flattened off towards the fixing location of the inlet blade(s) (20).
15. Compressor according to one of claims 9 to 14,
    if according to one of claims 8 to 10 additionally having a cylinder block and at least one, especially 5 to 9, piston(s) which is/are arranged to move axially back and forth in bores provided in the cylinder block,
    characterised in that
    the ratio of piston diameter and piston stroke (D/s) is approximately from 0.4 to 1.5, especially from 0.65 to 1.1, moreover especially approximately 0.95.
16. Compressor according to one of claims 9 to 15, especially claim 15,
    characterised in that
    the ratio of piston diameter and the throttling through-bore in the valve plate (D/d) is approximately from 1.5 to 5, especially from 2.5 to 4, moreover especially approximately 3.6.
17. Compressor according to one of claims 9 to 16, especially claim 15 or 16,
    characterised in that
    the ratio of the throttling through-bore in the valve plate and stroke of the inlet blade (d/t) is approximately from 2.5 to 8, especially from 3.7 to 6.7, moreover especially approximately 4.55.
18. Compressor according to one of claims 9 to 17, especially claim 15 or 16 or 17,
    characterised in that
    the ratio of piston stroke to the stroke of the inlet blade (s/t) is approximately from 10 to 30, especially from 14 to 24, moreover especially approximately 17.3.
19. Compressor according to one of claims 9 to 18,
    characterised in that
    the tilting behaviour of the tilt plate (1) acts in a self-limiting manner such that at high speeds of rotation of the compressor, especially at very high speeds of rotation or at the maximum speed of rotation, the angle of maximum deflection of the tilt plate (1) is less than the angle of maximum deflection α_max at low speeds of rotation of the compressor.
20. Compressor according to one of claims 9 to 19,
    characterised in that
    the geometry and dimensioning of all parts moved in translation, such as axial pistons, piston rod or sliding blocks or the like, on the one hand, and all parts moved in rotation, such as the tilt plate (1), members for conjoint movement or the like, on the other hand, are such that, for predetermined tilt angles of the tilt plate (1), especially between a predetermined minimum tilt angle and a predetermined maximum tilt angle, the moment M_k,ges due to the masses moved in translation, especially that of the pistons, where appropriate including sliding blocks, piston rods or the like, is so selected as to be less than the moment M_sw due to the moment of deviation, that is to say than the moment due to the mass inertia of the tilt plate, that at high speeds of rotation of the compressor, especially at very high speeds of rotation or at a maximum speed of rotation, the angle of maximum deflection α_max of the tilt plate (1) is less than the angle of maximum deflection at lower speeds of rotation of the compressor.

Images