DISPLACEMENT COMPRESSOR SYSTEM FOR R-718

Description

TECHNICAL FIELD

The refrigeration market is currently in flux and everyone is thus talking about the so-called “F-Gas Regulation” according to Regulation (EC) No. 842/2006 and No. 517/2014 on fluorinated greenhouse gases as a challenge, in order to roll back the predominant refrigerants HFC and HFO on account of their harmfulness for the climate and environment. There is thus a strong desire in the field of refrigeration technology for natural refrigerants, wherein in particular water stands out as a result of its good thermodynamic characteristics.

BACKGROUND

The widespread implementation of water as refrigerant R-718 (=water) has, however, foundered on the fact that, for example compared to ammonia in the same role, a displacement volume flow that is approximately 300 times greater is necessary for the same performance. As the pressure ratio that is ideally above a factor of 10 is quite high, the demands on a compressor increase tremendously, which must simultaneously still be oil-free and has to work as efficiently as possible under negative pressure, namely between 6 mbar and 200 mbar and, if necessary, still higher.

The disruptive character of water as a refrigerant is uncontested and will abruptly end the discussions conducted intensively world-wide regarding the known environmental and climatic problems with current refrigerants. Refrigeration technology can be represented via two broad areas here:

• • mobile refrigeration/air-conditioning technology (i.e. for trains, trucks and passenger cars)
  • stationary refrigeration technology (i.e. industrial refrigeration, commercial refrigeration and building air-conditioning, heat pumps)

Up to now, it has been attempted to meet this challenge with turbo-compressors, although these machines only generate lower pressure ratios of approximately 6 despite their two-stage configuration with intermediate cooling so that the necessary heat transfer at the condenser is realized merely to an unsatisfactory degree in the cooling circuit. In addition, there is with a turbomachine the serious disadvantage of its soft working characteristic (i.e. pressure values above volume flow) in order to be able to ensure stable operating points for various operating points.

A displacement machine for water-vapour compression is without question the better solution in order to overcome the challenges of water-vapour compression in R718 refrigeration circuits. For the compression of water vapour as refrigerant R-718, the following serious challenges must be resolved:

• • Conveyance of high water-vapour volume flows of well over 5.000 m³/h for, for instance, 35 kW refrigerating performance, which is 60 times greater than in the prior art. Consequently, new ground will have to be broken with respect to rotational speed as well as the geometrical design of the displacement machine.
  • Control of large pressure ratios of well over 10 in the event of low evaporator temperatures and higher temperatures in the condenser. As the isentropic exponent in the case of water vapour is simultaneously greater than 1.32 (present refrigerants have an exponent of approximately 1.1 and are thus hardly temperature-stressed), final compression temperatures of well over 200° C. result mathematically for refrigerants R-718, said temperatures only adversely affecting the effectiveness of the compressor, but also posing a risk in particular to the sensitive structural components of the compressor (especially the rotor mount on the outlet side).
  • The refrigerant circuit must be completely oil-free for R-718 (=water vapour), which represents a challenge in the case of twin-shaft displacement machines, because this type of machine requires, if an operating fluid (these days mostly oil) is omitted, a dry synchronisation for the spindle-rotor pair in order to avoid contact between the quickly rotating spindle rotors.

SUMMARY

The R718 displacement compressor system provides the following advantages:

(1) Safe avoidance of the consumption of play in the compressor:

• • The consumption of play in a compressor leads to its breakdown as a so-called “crash” by the contact of the working-space structural components, the gap values between the working-space structural components being reduced from the standard millimetre fractions to zero, even if not all over, but only where the mostly thermal expansions combined with rotational errors and other deviations among the variety of diverse influencing parameters can lead to the same. Such a consumption of play is to be avoided, by means of a sufficient safety margin for absolutely all operational, working and ambient conditions at all times in a reliable and complete manner.

(2) Best possible efficiency, i.e. optimum effectiveness for the R718 displacement compressor system:

• • Besides the operation with optimally adapted operational parameters, this relates in particular to the minimization of the system losses and in this context primarily the impact of inner gap leakages, which adversely affect efficiency, without jeopardizing the imperative goal of crash avoidance.

(3) Greatest reliability and high durability (long service life) of the R718 system:

• • In this context, it is primarily the sensitive structural components that need to be protected, in particular the rotor mount (especially on the outlet side) and the two drive motors with their respective equipment

(4) To the greatest possible extent *∘* independently of the external operating conditions in the sense that the R718 displacement compressor system adapts to the most varied conditions in an autonomous manner.

*∘* to the greatest possible extent in that there are practically no restrictions in said operating conditions.

(5) System intelligence:

• • The R718 displacement compressor system must be able, at all times and in all circumstances, to achieve the aforementioned advantages or, in the event of impending deviations, initiate in a timely manner corrective measures including the emission of external warnings and indications, in an autonomous manner by means of its own regulating mechanisms and regulating tools.

This object of compressing water vapour at pressures below atmospheric pressure by means of a twin-shaft rotary displacement machine is achieved in accordance with the disclosure by configuring the R718 displacement compressor system (42) as a closed vacuum system, comprising the core elements:

• • Evaporator (7) with evaporator housing vessel (29)
  • Condenser (8) with condenser housing vessel (28)
  • and compressor machine (41),
    wherein this compressor machine essentially comprises:
  • Compressor housing (1)
  • two autonomous rotary units (39 and 40) with electronic synchronization of the motor pair/spindle rotors via the FUs (2.4 and 3.4) by means of the FU control unit (16)
  • and bearing support unit(s) (25) on the side of the outlet for receiving the rotary bearing (4.2), and the outlet (12) with outlet openings via control balls (10) as well as end outlet openings (27)
    and everything is managed by the control unit (15).

The compressor machine (41) is configured here in accordance with the disclosure as an open machine that separates the evaporator (7) and the condenser (8) by means of the compressor housing (1). The compressor machine thus no longer has face-side compressor-limiting lateral parts (so-called “covers”) and can no longer be operated in an autarkic manner as a vacuum machine but rather only in conjunction with a connected evaporator (7) and a connected condenser (8) with respective housing vessels (28 and 29).

Moreover, the drive motors (2.3 and 3.3) for each spindle-rotor rotation unit (40 and 39) are located on the compressor inlet side (11) and suspended directly in the evaporator chamber (7) for optimal cooling of the electric drive motors (2.3 and 3.3) for the “unlimited” operation in accordance with the disclosure by purging the heat losses of these drive motors (2.3 and 3.3) via the respective refrigerant fluid flows K5 in the event of increased (i.e. above the nominal load) performance demands during operation. The monitoring of the drive motors (2.3 and 3.3) preferably occurs via temperature sensors in the area of the motor windings in order to be able to adapt the respective refrigerant fluid flows K5 accordingly so that the drive motors are not damaged and can provide the performance demanded.

Moreover, an intermediate water jacket (5), preferably with its own cooling tube coil (6), is part of this thermal balance management system according to the disclosure for the R718 displacement compressor system in order to operate the selective thermostatic control for the compressor housing (1) via the refrigerant fluid flow K1. While the selective interior cooling of the rotors (2.2 and 3.2) by means of a refrigerant fluid flow K2 and K3 (as 9.2 and 9.3) remains simultaneously possible, the situation with respect to the play between the working-space structural components (i.e. compressor housing and spindle-rotor pair) is regulated, and via the inner gap leakage thus the volumetric efficiency for the different operating/working points as a result of the respective thermal expansion behaviours of the structural components (which is saved in the control unit (15) for the regulation of the refrigerant fluid flows (9)), while crashes are simultaneously reliably avoided (crash as a disruptive consumption of play).

In order to implement the “unlimited” operation with simultaneous optimum efficiency and reliable crash avoidance, the thermal balance management according to the disclosure for the R718 displacement compressor system in the “maximum” version*∘* comprises the following refrigerant fluid flows (9) regulated in a selective manner via control unit (15):

• • 9.1 Refrigerant fluid flow (illustrated as K1) to the intermediate water jacket (5) via cooling tube coil (6)
  • 9.2 Refrigerant fluid flow (illustrated as K2) to the 2t rotor (2) via evaporator cooling bore (2.2)
  • 9.3 Refrigerant fluid flow (illustrated as K3) to the 3t rotor (3) via evaporator cooling bore (3.2)
  • 9.4 Refrigerant fluid flow (illustrated as K4) for injection cooling via centrifugal disk (22)
  • 9.5 Refrigerant fluid flow (illustrated as K5) for the cooling of each drive motor
  • *∘* Those applications pertain to the “maximum” version that have a particularly large scope of application, i.e. when the so-called “temperature lift” (as the difference between t_c and t_o) is greater than approximately 40 Kelvin, wherein smaller machines (i.e. with a displacement volume flow under approximately 5.000 m³/h at a nominal speed) are less sensitive here, i.e. higher “temperature lifts” are managed more smoothly.

This thermal balance management system is crucially necessary in particular for the working-space structural components during so-called k₀ operation, when the compressor is operated at a speed that, although creating the difference in pressure between the inlet and the outlet, does not yet convey or only conveys a minimum volume flow, as a result of which the compressor thus only contends with its own (interior) leakage, but would become accordingly hot on account of the power injection, which is reliably avoided by the thermal balance management system managed by the control unit (15) in accordance with the disclosure.

By means of this thermal balance management system which is practically independent of external conditions, the aforementioned “unlimited” operation is achieved in accordance with the disclosure, because the compressor machine (41) simultaneously achieves practically any pressure ratio and thus practically any pressure value with the corresponding temperature in the condenser in order to be able to purge the heat quantity in practically all surrounding conditions. This is the so-called “unlimited” operation, which does not exist in the prior art. Mentioned as an example here are the breakdowns of the air-conditioning systems in ICE trains that occurred because their air-conditioning systems could not handle the high outside temperatures. This can no longer happen with the solution in accordance with the disclosure.

Furthermore, a centrifugal disk (22) is preferably arranged on every spindle rotor on the gas inlet side (11) in accordance with the disclosure for the speed-optimized introduction of the injection refrigerant quantity K4, wherein the refrigerant-fluid-flow-K4 feed according to (23.1) or (23.2) on the preferably rough surface of the centrifugal disk ensures a speed-optimized refrigerant-fluid-flow-K4 mist, which is mixed with the gas flow in a sufficiently even manner. “Speed-optimized” is understood to mean here that the velocity vectors of the liquid refrigerant-fluid-flow-K4 mist droplets move similarly to the spindle-rotor surfaces, which is ensured by each centrifugal disk (22). A hard impact on the spindle-rotor surfaces as a result of large differences in velocity, and consequently at the least an unpleasant patter noise up to the damage of the spindle-rotor surfaces, is consequently reliably avoided.

In addition, there is a tooth flank offset Δk_vs(z) in the longitudinal rotor-axis direction z between the left and the right profile flank side so that the head angle be.2K(z) at the two-toothed spindle-rotor head arc becomes the be.2K.em(z) distribution in accordance with the disclosure, whereas the be.2K.stu(z) distribution results for selected distributions for μ.2(z) and μ.3(z) with a corresponding tooth height h(z) when there is no tooth flank offset. By representing the difference between be.2K.em(z) and be.2K.stu(z) as Δ.be.2K(z) distribution in conjunction with the pitch distribution m(z) and the distribution of the tooth heights h(z) in the longitudinal rotor-axis direction, it is evident that both the be.2K.em(z) distribution as well as the Δ.be.2K(z) distribution, put in a simplified manner, are configured as the opposite of the pitch distribution m(z):

If m(z) is at a maximum, both the be.2K.em(z) distribution as well as the Δ.be.2K(z) distribution are at a minimum. And if m(z) increases, both the be.2K.em(z) distribution as well as the Δn.be.2K(z) distribution decrease, whereas, in the event of a sharply decreasing m(z) in the inlet area, both the be.2K.em(z) distribution as well as the Δ.be.2K(z) distribution increase sharply. This feature according to the disclosure is valid with a precision of preferably ±15% and ensures on the inlet side (11), i.e. for the area of larger z values in the form of representation chosen here, higher volumes of the working chambers on the inlet side in order to consequently be able to suction larger displacement volume flows.

To this end, values above 0.6 (for example 10% to 15% higher) in the inlet area of the compressor machine (41) are also proposed for μ.3(z) so that the suction volumes increase further.

DESCRIPTION OF THE DRAWINGS

Furthermore, control balls (10) are provided for the selective adaptation of the inner compression ratios in accordance with the specific application, i.e. in particular in the event of different pressure values in the condenser as different working points during the operation of the R718 displacement compressor system. The inner volume ratio is, initially without taking thermodynamic effects into account, dependent on the geometry of the configured spindle-rotor pair as the simple ratio of the inlet working-chamber volume to the outlet working-chamber volume, which is determined at the time of manufacture of the spindle-rotor pairing. As various operating points with different pressure ratios (as outlet pressure p₂ divided by the inlet pressure p₁) are required, the control balls (10) ensure that efficiency-reducing overcompression is avoided in that the control ball is raised as a result of the pressure difference when the current outlet pressure p₂ is reached in the particular working chamber during compression, so that a partial gas flow leaves the working chamber in the direction of the outlet space (12) and thus to the condenser (8). This preferably occurs both in the longitudinal direction of the rotor axis as well as on the face side at the outlet end (12) in accordance with the illustrative representation represented in FIG. 2. The control balls (10), which are preferably weight-loaded, are raised by the difference in pressure between the current pressure in the particular working chamber and the outlet pressure p₂ and move back by the force of gravity, which is shown by means of the angles in accordance with the illustrative representation in FIG. 8, wherein g indicates the direction of gravity. Alternatively, it is of course also possible to implement a simple spring engagement for the control balls.

The control balls (10) in FIG. 1 and in FIG. 4 can be seen on the outlet side in the outlet control disk (12) as well as in FIG. 2 in the longitudinal direction of the rotor axis for the Π_iV adjustment for avoiding efficiency-damaging over- and undercompression as well as in an axial top view in the outlet control disk (12) and adapt the so-called inner volume ratio Π_iV in accordance with the specific application to the particular pressure ratio actually present.

Moreover, an intermediate support (17) on the two-toothed spindle rotor (2) is proposed for weight reduction, in particular also as a lower mass moment of inertia during initial acceleration (as well as deceleration) with simultaneous high flexural rigidity, for example made of a vacuum-compatible fibre-composite material, e.g. as a CFRP material.

Furthermore, there will generally be different application scenarios with various temperature-lift application ranges with various “volume curves” (i.e. the distribution of the working-chamber volumes between inlet and outlet in the longitudinal direction of the rotor axis) as various application-specific requirements so that various spindle-rotor pair designs in particular with respect to an energy-efficient mode of operation are also advantageous and useful. In order to avoid having to completely configure each compressor machine individually, it is proposed in accordance with the disclosure that various spindle-rotor pairings can be inserted in the practically identical*∘*∘ compressor housing shown illustratively in FIG. 13. This is designated as “rotor construction kit” and means that, for various application scenarios, the individually most efficient adaptation to the particular user requirements is achieved in an uncomplicated manner by means of a simple and direct changing of the spindle-rotor pair.

• • *∘* By “practically identical” it is understood that, for example, the crossing angle α (shown in in FIG. 13) remains the same, while the rotor length and possibly also the diametric distribution in the longitudinal direction of the rotor axis can vary while the compressor housing sleeve remains identical, the simple and rapid implementation of various volume curves via different spindle-rotor pairs being the core function of the “rotor construction kit”. The simple and flexible manufacture of the spindle rotors (generally by turning) is an important feature here for a simple realization.

The inner volume ratio (i.e. the simple quotient of the working-chamber volume at the inlet divided by the working-chamber volume at the outlet) of the spindle-rotor pair is limited to an iV range preferably between 2 up to a maximum of 20, wherein the adaptation to the particular working/operating point with its current actual pressure ratios occurs via the aforementioned control balls (10) in accordance with the specific application. If still greater temperature lifts ΔT_h in accordance with

ΔT_h=t_c−t₀

with correspondingly higher pressure ratios are necessary, mostly momentarily, during operation, a so-called undercompression occurs (the pressure of the last working chamber is lower than the pressure at the outlet) and the last working chamber is pushed out in an isochoric manner against a higher pressure at the outlet (12). In order to curb this process which reduces the efficiency of the compressor, it is proposed in accordance with the disclosure that the play values in the compressor outlet area are selectively increased by approximately 20 to at least 50% greater average gap clearances, preferably realized simply in that the outer rotor diameters are manufactured to be correspondingly smaller over an area in the longitudinal direction of the rotor axis corresponding to 0.3 to 2 times the extension of the working-chamber length on the outlet side in the longitudinal direction of the rotor axis, wherein, in the event of a face-side outlet plate with a bearing support (25) on the control edge (27.S), this is also realized by bevelling (in the sense of rendering oblique) said control edge (27.S).

These measures in accordance with the disclosure combined with the simultaneous limitation of the inner-volume ratio range at the spindle-rotor pair, preferably to the aforementioned iV range, are called

• • “outlet-gap-iV adaptation”.

The outer rotor diameter/gap adaptation on the rotor pair preferably occurs here so that this diametric adaptation, which progresses in the direction of the outlet initially slowly, increases to progressively larger values so that the averaged gap clearances reach the above-mentioned increase as an average. This outlet-gap-iV adaptation helps in particular to reduce noise as the pressure pulsations on the outlet side are dampened.

For good measure, the “PIRSA” procedure is proposed for the R718 displacement compressor system (42) with its respective spindle-rotor pairs, preferably for every working/operating point: “PIRSA” stands for “Pressure/Inner Ratio/Speed Adaptation”. It is known that various working/operating points can be realized by means of different operating parameters (mentioned illustratively in the following). By means of “PIRSA”, the operating parameters are adjusted via the control unit (15) so that the power input for the R718 displacement compressor system (42) is minimal for the particular working/operating point required in accordance with the specific application.

As operating parameters, this is especially valid regarding:

• • Regulation of the refrigerant fluid flows (9) in particular with regard to the injection quantity (9.4)
  • Adaptation of the spindle-rotor speeds via the CU-FU (16) for a certain suction capacity
  • Setting of the pressure values in the evaporator (8) and in the condenser (9)

The control unit (15) has its own preinstalled databank here and can adapt these operating parameters in a regulating manner, wherein this process occurs through self-learning by means of trial and error in accordance with the specific application, by modifying individual values slightly and determining by the reaction of the system whether the overall efficiency improved or suffered. This way, the databank is constantly broadened in every operating point through self-learning and the system becomes increasingly more intelligent in terms of efficiency improvement.

Brief explanation regarding the tooth flank offset ΔkVs(z) for each profile flank side:

or every rotational angle position φ, there is, corresponding to the transmission ratio for each spindle rotor, a z-position as a z(φ)-function, the derivation of which via the equation below then yields the so-called pitch distribution m(φ) for each spindle rotor, wherein a distinction is additionally made in accordance with the disclosure between the right and the left profile flank side via the index s:

m s  ( ϕ ) = 2  π · dz s  ( ϕ ) d   ϕ ≈ 2  π · Δ   z s  ( ϕ ) Δ   ϕ

As the distinction between the right and left profile flank side is difficult and often leads to confusion with regard to the perspective as well as its dependence on the pitch direction (i.e. right- or left-handed) for each rotor, the tooth flank offset according to the disclosure is illustrated via the head arc angle be.2K(z) in accordance with FIG. 9 in a simplified manner in the plane, although this type of problem is three-dimensional due to the non-parallel rotational axes of the spindle rotors.

Brief explanation regarding the formation of the tooth profile: (simplified as a plane representation)

The various tooth heights h(z) in the longitudinal direction of the rotor axis (generally designated by z) are generated via the so-called μ values at each rotor, as the following equations are valid for the tip radii for each spindle rotor:

on the 2t rotor:

R_2K(z)=μ₂(z)·a(z)  (Eq. 1.1)

on the 3rotor:

R_3K(z)=μ₃(z)·a(z)  (Eq. 1.2)

Accordingly, the following equation is valid in the longitudinal rotor-axis direction z for the tooth height h(z):

h(z)=(μ₂(z)+μ₃(z)−1)·a(z)  (Eq. 1.3)

The distributions for μ.2(z) and μ.3(z) are preferably chosen so that the requirements of the specific application are fulfilled to the highest possible degree, for example with respect to working-chamber volume as well as the so-called “volume curve” (i.e. the distribution of the working-chamber volumes in the longitudinal direction of the rotor axis, wherein in particular the variation of these working-chamber volumes is of importance). The following holds for μ.3(z) here:

• • For, μ₃(z)≤0.6 the pairing of the 2rotor (2) and the 3t rotor (3) remains without a blowhole.
  • In order to increase the working-chamber volume on the inlet side, it can be useful for some applications to increase the μ.3(z) value above this value of 0.6.
  • The μ.2(z) value can be chosen freely, although, besides the tooth heights h(z), the remaining base circle thicknesses are of importance, with the objective that the critical bending speeds for each spindle rotor are realized in accordance with:

ω kritisch 2 · Rotor = 1 , 5 · ω kritisch 2 · Rotor Critical 2 · rotor Critical 3 · rotor ω kritisch allgemein = c m Generall y   critical   with   ( simplified )

• •  In particular via the values for μ.2(z) and μ.3(z) as well as regarding the rotor length and the rotor masses, the critical bending speed of the faster-rotating 2t rotor (3) is realized 1.5 times higher than with the 3t rotor (3).

With these points, the aforementioned advantages are achieved by way of the present disclosure:

The thermal balances of the working-space structural components, i.e. the housing (1) and the spindle-rotor pair (2 and 3), in the R718 displacement compressor system (42) are managed and regulated so that the following advantages are simultaneously met at all times and in all conditions and intelligently by the system:

• • (1) Safe avoidance of play consumption (so-called “crash” by contact of the working-space structural components) due to the fact that, for the different operating/working points, the various thermal expansions of the working-space structural components work with measured reference-temperature values by means of the intelligent management of the thermal balances with expansion behaviours of the working-space structural components saved in the control unit (15) for the different temperature levels at each operating/working point.
  • (2) Minimization of inner gap leakage by observance of a gap range, preferably in a range of ±25%, wherein the lower value is derived from the avoidance of gap consumption plus a safety margin and lies in the range of 0.05 to 0.1 mm with a corresponding rotational precision of preferably less than 0.02 mm for machine sizes with a range of axle-separation distances of approximately 100 mm to approximately 500 mm (below this, the value is correspondingly smaller; above, larger)
  • (3) Protection of the sensitive structural components, in particular the rotor mount (especially on the outlet side) and the two drive motors by means of the described purge-gas system (30 and 31)
  • (4) To the greatest possible extent*∘* independently of external conditions of application in the sense that the R718 displacement compressor system adapts to the most varied conditions in an autonomous manner. *∘* to the greatest possible extent so that there are practically no restrictions regarding conditions of application.

(5) Intelligent management via the control unit (15), in particular of the refrigerant fluid flows as well as in accordance with PIRSA so that the R718 displacement compressor system has the respectively lowest energy requirement in every operating point, i.e. works with the greatest efficiency and simultaneously achieves the aforementioned advantages.

The necessary capability for accomplishing these advantages in accordance with the disclosure in the sense of intelligence lies in the control unit (15). Both its design as well as its operation must be configured in accordance with the disclosure so that the advantages mentioned in the introduction are reliably achieved at all times.

In order to meet these advantages, the following regulating variables are available:

• • K1. Housing thermal-balance management
  • K2. 2t spindle-rotor thermal-balance management
  • K3. 3t spindle-rotor thermal-balance management
  • K4. Injection for evaporation cooling during the compression process
  • K5. Cooling of the motors

The speed adaptation occurs via FUs (2.4 and 3.4) via the electronic synchronization of the motor pair/spindle rotors by means of the FU-CU (16) in conjunction with the control unit (15).

CET stands for Compressor End Temperature=i.e. the temperature at the gas outlet of the compressor

The injection cooling K4 performs the main share of cooling during compression, whereas the cooling of the working-space structural components is added by the control unit (15) in particular to compensate for various thermal expansions of each working-space structural component and/or to protect the sensitive structural components (in particular the rotor mount as well as the drive motors) by saving this in the algorithm of the control unit (15).

Evaporator (7) with the (lower) pressure p₁ and the temperature t₀ before the displacement compressor machine

Condenser (8) with the (higher) pressure p₂ and the temperature t_c after the displacement compressor machine, which compresses the refrigerant R-718 from p₁ to p₂, wherein the refrigerant R-718 undergoes the temperature increase from t₀ to t_c.

Fundamental Explanation

The cooling water “Kü” generally purges the heat Q_ab from the condenser (8), while the heat Q_ent is withdrawn from the chilled water “Ka” in the evaporator (7) by the displacement compressor system.

Designated as the refrigerant (abbreviated as “K”) here is the water that is diverted from the evaporator (7) as a refrigerant fluid flow in a manner regulated by the control unit (15) in the refrigerant separator (26) for separation into a main flow HS and the individual refrigerant fluid flows K1, K2, K3, K4 and K5 for the achievement of the aforementioned advantages.

(1) The heat balance for the compressor housing (1) is selectively set in accordance with the disclosure as the so-called “housing thermal-balance management” via the intermediate water jacket (5) by the control unit (15) in accordance with the specific application as set out below:

• • a) cooled via external cooling water “Kü” in the intermediate water jacket (5) when the control unit (15) determines from the actually present temperature values (in particular for “Kü”) in comparison with those stored in the databank in the control unit (15) that the available cooling water temperatures in accordance with the specific application are favourable (in most cases in the sense of low enough) for the housing thermal balance in order to institute play settings between the compressor housing (1) and the spindle-rotor pair (2 and 3) which, first, avoid a crash (as play consumption) reliably while, second, ensuring the optimum efficiency of the compression with respect to inner gap leakage, specifically: The gap values for crash avoidance lie in the range of 0.03 to 0.05 mm, a safety margin (for example because of rotational deviations) of approximately 30% to 50% being added so that the lower gap values result, designated as ΔSp.u. The upper gap values ΔSp.o should preferably not be greater than ΔSp.u by a factor higher than 2.
  •  By means of the various thermal expansion behaviours of the working-space structural components (i.e. essentially the housing and the rotor pair), the so-called thermal-balance management system now has to maintain, via the refrigerant fluid flows (9) regulated by the control unit (15), the actual gap values between ΔSp.0 and ΔSp.o in accordance with the specific application.
  • b) If the cooling water temperatures available in accordance with the specific application are unfavourable (in most cases in the sense of too high) for the housing thermal balance, then the control unit (15) ensures via the refrigerant fluid flow 9.1 (illustrated as K1), for example by means of a regulation organ (26) and simple cooling pipe coil (6) in particular in the outlet area, that the rising heat in the intermediate water jacket (5) is purged, wherein rising in the sense of the extension in the longitudinal rotor-axis direction depending on the convection in the intermediate water jacket as the intermediate medium carrier and compensation of temperature differences that are too high.

(2) unlimited through internal cooling during operation independently of external conditions and self-adjusting, i.e. at 5° C. ambient conditions as well as at 60° C.=indication of limits no longer necessary=the condenser temperature is automatically increased and the inner cooling adapts automatically, i.e. no more requirements regarding max. admissible cooling water temperature=in accordance with the disclosure, everything is now feasible

(3) in particular the e-motors can be overloaded practically at will thanks to the adaptable intensive cooling

(4) Intermediate water jacket (5) on the compressor housing (1) with insulation jacket (20) toward the condenser (8)

(5) Compressor with open inlet (11) and outlet (12), there are no longer any lateral housing end parts (“covers”), it is no longer a classically autarkic compressor, but rather an open machine

(6) Cooling mechanisms diverted and distributed by the control unit as so-called “working-space structural-component thermal management system”:

• • HS is the main flow for achieving the basic object between heat absorption in the evaporator and the release of heat in the condenser
  • K1 Cooling for the compressor housing, preferably as evaporator cooling via intermediate water jacket (also only so much that the resulting gap values via the minimized rotor cooling K2 and K remain in a selected range, e.g. preferably within ±25%)
  • K2 Structural-component cooling for the two-toothed spindle rotor→minimized(!) primarily for the protection of the bearings
  • K3 Structural-component cooling for the three-toothed spindle rotor→minimized(!) primarily for the protection of the bearings
  • K4 Cooling by refrigerant injection→bears the brunt as the most important variable, i.e. >80%
  • K5 Cooling for each drive motor→only for the maintenance of operation (monitored & managed by the motor thermal elements, preferably in the motor coils)

(7) Structural-component cooling K1 and K2 and K3 for the implementation of two main requirements:

• • Command of the play settings in order to be able to compensate for various thermal expansions, wherein the play values should preferably remain within approximately ±25%.
  • command of k₀ operation as well as minimum displacement volume flows in a reliable and sustained manner
  • Avoidance of temperatures that are too high for critical structural components, in particular bearings on the outlet side

(8) Injection K4 as the main cooling mechanism by means of evaporation during the compression

• • Objective: fine mist as largest possible surface for efficient evaporation as heat transfer during compression
  • Centrifugal disks with rough surface and terminal inclination for the avoidance of streamlets, for a distribution as even as possible
  • Centrifugal disk with outer groove, if appropriate with radial drainage bores, in order to reduce slippage
  • Feed to the centrifugal disk or, if appropriate, as drainage bore in the bottom via double tube or via the support arm for the suction bearing support
  • Injection instead of centrifugal disk via bores in the bottom at the inlet (unlikely)
  • Use of injection as regulation of the actual inner compression: The evaporating liquid causes a sharp volume increase in the working chamber with corresponding increase in pressure

(9) The distance between the spindle-rotor axes at the inlet (11) preferably at least 10% greater than at the outlet (12)

(10) Adaptation of the inner volume ratio Π_iV via vacuum-compatible control balls (10), which are preferably weight-loaded and pushed aside by the difference in gas pressure and also return by the force of gravity to a (preferably elastomer) ramp (10.R) inclined at the angle γ_R when Δρ falls again, configured

• • over the rotor length (represented illustratively in FIG. 2)
  • as well as on the end plate (12) as control disk (represented illustratively in FIG. 1)

(11) Outlet end plate as control disk (12) via peeling disks for the ideal play adjustment for the face-side gap between the end of the rotor and the end plate individually for each spindle rotor

(12) The effort for the iV adaptation (e.g. via control balls) can be drastically reduced in accordance with the specific application by the adjustment of the respective pressure values both at the condenser as well as at the evaporator with a simultaneous volume flow adaptation so that the pressure ratio of these two pressure values corresponds to the inner volume ratio Π_iV of the compressor so that an over- or undercompression is kept within acceptable limits avoided in accordance with PIRSA=Pressure/Inner Ratio/Speed Adaptation

(13) Pitch distribution via the tooth flank offset Δk_vs(z) varied between the right and left tooth flank for the maximization of the cross-sectional surface area in each end section in particular in the suction area: As the right tooth flank in a 2t spindle rotor configured as right-handed has the distribution vis-à-vis the left tooth flank that is represented illustratively, the tooth width of the 2t spindle rotor is reduced for the purpose of maximizing the traverse-section working-chamber scooping surface areas in the suction area designated as tooth flank offset of the flanks in relation to one another in accordance with the disclosure

(14) cylindrical inner cooling of spindle rotor can be limited to the last area, i.e. not over the entire rotor length (with corresponding increase in the bottom wall thickness at the inlet)

(15) the maximum version is represented (so to speak the “Mercedes”), as all cooling mechanisms are realized—there will also be a slimmed-down version (so to speak the “VW”), by preferably/for example omitting the structural-component cooling and adjusting the temperatures during compression only via the injection cooling, i.e.: the above advantages can only be achieved in a limited manner, as this is sufficient for several applications.

(16) Drive motors on the inlet side (on account of constructional space as well as temperature protection with overload option)

(17) K₀ speed measurement (as self-diagnosis for the determination of changes, e.g. formation of deposits, etc.)

(18) CFRP intermediate support (17) on the two-toothed spindle rotor for weight reduction, in particular the mass moment of inertia when starting (accelerating) with simultaneously high flexural rigidity

(19) Configuration for purging via the intermediate spaces by means of a bypass bore for each rotor mount

(20) Circumferential overflow groove in the evaporator and drain at the deepest point for the operating modes according to FIG. 14 as an alternative to the closed circuit with the riser pipe (19)

(21) Mixing tap and mixing section as option for selective temperature adjustment

(22) preferably with CO2 cascade system for lower temperatures

FIG. 1 illustratively shows a representation of a longitudinal section through the R718 displacement compressor system (42) in accordance with the disclosure with a standing configuration. The compressor machine (41) separates the evaporator (7) with the lower values for pressure p₁=p₀ and temperature t₀ from the condenser (8) with the higher values for pressure p₂=p_c and temperature t_c, each with a surrounding, vacuum-maintaining housing vessel (28 and 29), preferably cylindrical and fixed on the compressor-housing extension (1.P) in accordance with FIG. 5. The customary vacuum pump for ensuring negative pressure is not shown in the represented R718 displacement compressor system (42), but is sufficiently known and implemented in accordance with FIG. 4 when purging preferably via the shielding-gas discharge (31) shown in FIG. 4.

For a more detailed illustration, FIG. 3 shows an enlargement of the inlet area with the evaporator (7) of this representation and FIG. 4 shows an enlargement of the outlet area with the condenser (8). The important gap values between the rotor head and the housing are adjusted by means of peeling disks (18) on the inlet side via the positions of the spindle-rotor units (39 and 40 in accordance with FIG. 13) in the longitudinal rotor-axis direction. Likewise, the face-side gap values of the spindle rotor to the outlet control disk (12) are adjusted via peeling disks (18).

FIG. 2: As an illustrative sectional representation relating to FIG. 1 perpendicular to the axis of the housing vessel (28) approximately halfway down the longitudinal rotor axis in a perspective looking toward the outlet (12) as a cylindrical cooling-system cross section for the R718-displacement compressor system with control balls (10) both in the longitudinal rotor-axis direction (the control balls are accordingly shown as sectioned and shaded) and as a top view of the outlet control disk (12) simply as circular control-ball openings, which are shown in turn in FIG. 1 as sectioned and shaded at the outlet (12). Furthermore, the end outlet openings (27) with the control edges (27.S) can be seen clearly on the outlet control disk (12).

FIG. 3: As an illustrative representation for the suction area shown in FIG. 1 with a representation of the respective refrigerant fluid flows K1 and K2 and K3 and K4 and K5 with HS as the circuit-medium-R718 main flow for the achievement of the core objective for the transfer of heat. Besides the different possibilities for the refrigerant-fluid-flow feed, various configurations are shown simultaneously (in practice they are realized separately) for the heat transfer in the evaporator (7), by having the refrigerant R-718 flow, for example, via the overflow groove (37) over a large enough surface to the drain (37.a) OR via separate heat-exchanger surfaces (38), for example as an embedded heat-exchanger pipe system (38) on the floor of the housing vessel (29) in order to ensure the represented heat transfer Q_ent from the evaporation process. In order to minimize undesired heat transfers here, different insulation approaches (29.i) are represented, e.g. via an insulation layer (left side in FIG. 3) or via an evacuated intermediate space (right side in FIG. 3).

FIG. 4: As an illustrative representation for the outlet area shown in FIG. 1 with a representation of the housing cooling system via the intermediate water jacket (5) with cooling pipe coil (6), moreover control balls (10) in the outlet control disk (12) and a purge-gas apparatus for each shielding-gas feed (30) and discharge (31) via a buffer space (13) via the aforementioned vacuum pump creating suction for the desired negative pressure in the displacement compressor system, moreover illustratively with the regulating organ (26) for the separation of the refrigerant fluid flows K1=1, K2=2, K3=3, K4=4, and K5=5 (designated as “KM-TS.Δ”) as well as HS=0 as circuit-medium-R718 main flow (as “KM-HS”), said separation being regulated by the control unit (15) in accordance with the specific application.

FIG. 5: As illustrative 3D representation toward the compressor housing (1) with separation between evaporator space and condenser space via the preferably cylindrical dividing plate (1.P), moreover with the preferably three attached bearing-support support arms (24) for each bearing support (25). In this representation, the principle of the “open” compressor becomes clear, as merely the two spindle-rotor rotation units (39 and 40) are mounted in the housing, and the compressor machine (41) is practically complete without specific face-side closing parts (thus “open” compressor)

FIG. 6: Illustratively represented is an enlargement of the feed (23.1) of injection refrigerant K4 to the centrifugal disk (22) via a bearing-support support arm (24) in the compressor inlet area (11) shown in FIG. 1, by having the refrigerant fluid flow K4 reach, for example via a small tube, the upper side of the centrifugal disk (22), where K4, which is distributed evenly by means of centrifugal forces, then enters the gas flow at the inlet (11) with the optimum speed profile vis-à-vis the spindle rotor.

FIG. 7: As an illustrative representation of the centrifugal disk (22) preferably with rough, course surface for the reduction of slippage and better distribution of the refrigerant fluid flow K4, wherein the diameter øa.s as well as the height h and the angle γ_s substantially influence the spray of the refrigerant fluid flow K4 from the centrifugal disk and are to be configured in accordance with the specific application.

FIG. 8: As an illustrative representation of the control ball (10) for the adaptation of the inner volume ratios to various temperature lifts with corresponding pressure differences, rolling away on the ramp (10.R) via the angles γ_A and γ_R with respect to the direction of gravity g while preferably subject to the force of gravity by means of the difference in pressure Δ_p at the control ball between the particular working-chamber pressure and the outlet pressure, a special material not necessarily being necessary here, and automatically rolling back by force of gravity when the difference in pressure decreases.

FIG. 9: Illustrative representation of the spindle-rotor profile pairing, wherein the problem, which is actually three-dimensional on account of non-parallel rotational axes, is shown in a simplified manner in a plane. As a section through the end of the spindle-rotor pair showing the head arc angle be.2K(z), which yields the tooth flank offset Δk_vs(z) as a different z(φ) distribution for the right and left profile flank for each tooth, as well as with the particular μ value for each spindle rotor via the rotor head circles with a(z) distribution for the distance between the rotor axles.

FIG. 10: When determining the rotor profile pairing, the following functional distributions, illustratively for the crossing angle α=15° between the rotational axes of the rotor at a rotor profile length of L=376 mm, are shown, which portray the tooth flank offset Δk_vs(z) via the Δ.be.2K(z) distribution in relation to the pitch distribution m(z) in the longitudinal rotor-axis direction z, wherein this z progression is always chosen as the abscissa, i.e. the value range: 0≤z≤L. The outlet (12) is at z=0 and the inlet (11) is at z=L here

FIG. 10.1: [Values Only Illustrative]

The correlation between the rotational-angle extension parameter φ for the range 0°≤φ≤1320° and the z position as a z(φ) function yields the pitch distribution m(φ) via the known equation:

m  ( ϕ ) = 2  π · dz  ( ϕ ) d   ϕ

Applied over the longitudinal rotor-axis direction, the represented pitch distribution m(z) then results, which begins at z=0 mm with 28 mm, then quickly increases to a pronounced maximum range before the inlet (11), the pitch falling at z=L quickly back to 78 mm.

FIG. 10.2: [Values Only Illustrative]

The tooth height h(z) in the longitudinal rotor-axis direction results for each axis separation-distance value a(z) in accordance with the crossing angle via the meshing spindle-rotor heads, wherein the rotor-head radius values then result via the respective μ values from the following equations:

R_2K(z)=μ₂(z)·a(z)

and

R_3K(z)=μ₃(z)·a(z)

and

h(z)=(μ₂(z)+μ₃(z)−1)·a(z)

The μ values shown in FIG. 10.2 lead here to the cylindrical rotor mount shown in FIG. 1 in order to make possible the evaporator cooling at each spindle rotor in particular during k₀ operation.

FIG. 10.3: [Values Only Illustrative]

As a continuation of FIG. 10.2, two configurations relating to the μ values shown for the head arc angle be.2K(z) on the two-toothed spindle rotor are represented:

• • In cases where there is no tooth flank offset between the right and left profile flank side, the dashed-dotted be.2K.stu(z) distribution results for the head arc angle, the distribution falling continuously from 65° at z=0 to 53° at z=L.
  • The be.2K.em(z) distribution in accordance with the disclosure starts at z=0 for example also at 65°, then exhibits a very different distribution, as there is a minimum in the last third of the rotor length on the inlet side, after which it then increases rapidly to 70° at z=L.

FIG. 10.4: [Values Only Illustrative]

As the continuation of FIG. 10.3, instead of the p distributions, the pitch m(z) in accordance with FIG. 10.1 is now depicted and the opposite distribution between be.2K.em(z) and m(z) is clearly visible

FIG. 10.5: [Values Only Illustrative]

Complementing FIGS. 10.3 and 10.4, the difference between be.2K.em(z) and be.2K.stu(z) as the Δ.be.2K(z) distribution is additionally depicted in conjunction with the pitch distribution m(z) so that the idea in accordance with the disclosure with regard to the tooth flank offset Δk_vs(z) between the left and the right profile flank side becomes clear.

FIG. 11: As an illustrative representation, three different spindle-rotor pairs for the rotor construction kit are represented as FIG. 11.1 and FIG. 11.2 and FIG. 11.3, which fit with respect to their outer/connecting geometry to the same compressor housing (1), at the least to the same housing sleeve, wherein the following description applies:

FIG. 11.1 shows illustratively a spindle-rotor pair with a high suction capacity with a moderate number of tiers for applications in which it is less the compression capacity that is of importance, but a high volume flow.

FIG. 11.2 shows illustratively a spindle-rotor pair with a moderate suction capacity with an intermediate number of tiers for applications without a pronounced prioritization, i.e. more of a general orientation.

FIG. 11.3 shows illustratively the spindle-rotor pair with a low suction capacity with a very high number of tiers for applications in which a high compression capacity is more important than volume flow.

FIG. 12.1: Illustrative representation of an alternative design for the outlet control disk (12), in which control balls (10) can be omitted, by configuring the outlet control disk to be rotatable (12.S) together with the pivot bearings (12.g) as well as the discharge slot (12.s) at the end of the two-toothed spindle rotor.

FIG. 12.2: Illustrative representation of the configuration of the rotatable outlet control disk (12.d) with the discharge slot (12.s) as well as, in addition, lateral outlet notches (12.k), so that the last working chamber pushing outward does not close again when the minimum inner compression ratio is adjusted. These notches are necessary because the discharge slot (12.s) cannot take up too much of the circle as otherwise the factor by which the inner compression can be increased sinks. When the maximum inner compression ratio is set, the outlet notches (12.k) are above the chamber that would open without the outlet control disk (12.d). The outlet notches (12.k) are then closed laterally by the outlet plate, which closes off the 3t spindle rotor, and the compressor housing (1). Shortly before the last working chamber has completely emptied, the compressor housing (1) is laterally removed and the 3t spindle rotor is open.

FIG. 13: Represented illustratively are the finished and completely balanced spindle-rotor rotation units (39 and 40), which, without further intervention via the peeling disks (18), can be inserted without any changes in the compressor housing (1) for the exact gap adjustment and thus form the open compressor machine (41).

FIG. 14: Represented illustratively are 3 operating modes for operating the R718 displacement compressor system for different cooling-water (“ü”) and chilled-water (“a”) temperature levels, which are optimally set in accordance with the specific application by the control unit (15) by means of PIRSA as “Pressure/Inner Ratio/Speed Adaptation”, as described.

Terms such as substantially, preferably, and the like as well as possibly, as indications of imprecision, are to be understood in the sense that a deviation of plus/minus 5%, preferably plus/minus 2% and especially plus/minus one percent from the standard value is possible. The applicant reserves the right to combine any features and also any sub-features from the claims and/or any features and also partial features from a sentence of the description in any way with other features, sub-features or partial features, also beyond the features of independent claims.

In the different figures, parts that are equivalent with respect to their function are always provided with the same references so that these are generally only described once.

In the displacement system for the refrigerant R718 with a compressor machine (41), an evaporator (35) and a condenser (36), the open compressor machine (41) is configured as a spindle-rotor compressor in the form of a twin-shaft rotary displacement machine for conveying and compressing gaseous media. It has a spindle-rotor pair (2 and 3), which is arranged in a compressor housing (1) and configured with an electronic synchronization of the motor pair/spindle rotors. The compressor machine (41) is arranged between the evaporator (35) and the condenser (36).

Terms such as substantially, preferably, and the like as well as possibly, as indications of imprecision, are to be understood in the sense that a deviation of plus/minus 5%, preferably plus/minus 2% and especially plus/minus one percent from the standard value is possible. The applicant reserves the right to combine any features and also any sub-features from the claims and/or any features and also partial features from a sentence of the description in any way with other features, sub-features or partial features, also beyond the features of independent claims.

In the different figures, parts that are equivalent with respect to their function are always provided with the same references so that these are generally only described once.

REFERENCE LIST

• • 1. Compressor housing preferably simultaneously with ø-dividing plate (1.P) between evaporator (7) and condenser (8) with a distance between the spindle-rotor-receiving bores that is at least 15% greater on the inlet side than on the outlet side, wherein these bore axes are preferably configured so as to intersect (i.e. with perpendicularity) or cross (i.e. skewed).
  • 2. Spindle rotor, preferably with two-toothed gas-displacing outer thread, called “2t rotor” for short, preferably comprising an aluminium alloy with good thermal conductivity (preferably over 150 W/m/K), rotationally fixed via brace points on a steel shaft (2.1) with a cylindrical evaporator cooling bore (2.2) oriented inward and driven directly on the inlet side by its own drive motor (2.3) controlled by its own frequency converter (2.4), designated as “FU.2”, and by an FU control unit as “FU-CU” (16) via electronic motor-pair/spindle-rotor synchronization
  • 3. Spindle rotor, preferably with three-toothed gas-displacing outer thread, called “3t rotor” for short, preferably comprising an aluminium alloy with good thermal conductivity (preferably over 150 W/m/K), rotationally fixed via brace points on a steel shaft (3.1) with a cylindrical evaporator cooling bore (3.2) oriented inward and driven directly on the inlet side by its own drive motor (3.3) controlled by its own frequency converter (3.4), designated as “FU.3”, and by an FU control unit as “FU-CU” (16) via electronic motor-pair/spindle-rotor synchronization
  • 4. Mount for each spindle rotor, which is mounted at both ends, preferably configured on the inlet side as a fixed bearing (4.1) for axial and radial forces and on the outlet side as a floating bearing (4.2) which is preferably engaged in a cushioning manner
  • 5. Intermediate water jacket for regulation of the thermal balance for the compressor housing (1) with external thermal insulation (20) toward the surrounding condenser (8)
  • 6. Cooling pipe coil in the intermediate water jacket, which is preferably guided so as to be closer together on the outlet side and ends toward the evaporator (7)
  • 7. Evaporator with the (lower) pressure p₁−p₀ and the temperature t₀ before the compressor machine (41)
  • 8. Condenser with the (higher) pressure p₂=p_c and the temperature t_c after the compressor machine (41), which compresses the refrigerant R-718 from the pressure p₁ to p₂, wherein the refrigerant R-718 fundamentally undergoes the temperature increase from t₀ to t_c.
  • 9. Respective refrigerant fluid flows for complete management, regulated by the control unit (15), of the thermal balances of the working-space structural components, i.e. housing and rotor pair, as well as for the compression process as defined below:
  • 9.1 Refrigerant fluid flow (illustrated as K1) to the intermediate water jacket (5) via cooling pipe coil (6)
  • 9.2 Refrigerant fluid flow (illustrated as K2) to the 2t rotor (2) via evaporator cooling bore (2.2)
  • 9.3 Refrigerant fluid flow (illustrated as K3) to the 3t rotor (3) via evaporator cooling bore (3.2)
  • 9.4 Refrigerant fluid flow (illustrated as K4) for injection cooling via centrifugal disk (22)
  • 9.5 Refrigerant fluid flow (illustrated as K5) for the cooling of each drive motor
  • 9.6 The circuit medium-R718 main flow for achieving the core object for the heat transfer (e.g. as heat pump, or in the refrigeration technology process) is represented as HS.
  • 10. Control balls that are vacuum-compatible for the adaptation of the inner volume ratio Π_iV for various working points for the avoidance of an efficiency-reducing over- or undercompression both in the longitudinal rotor-axis direction as well as for each control disk (12) in accordance with the desired field of application with a ramp (10.R) inclined at an angle γ_R vis-à-vis the direction of gravity “g” in accordance with FIG. 8
  • 11. Gas inlet as open collection space for the medium with the gas pressure p₀ (for simplification, pressure losses in the lines are neglected for the time being)
  • 12. Gas outlet as a control disk for each spindle rotor with defined outlet openings for the medium at the gas pressure p_c (for simplification, pressure losses in the lines are neglected for the time being)
  • 13. Neutral collection/buffer space for each working-space shaft passage with reduced gas pressure vis-à-vis the system pressure, preferably created e.g. by means of a negative-pressure/vacuum pump in order, as a “purge-suction chamber”, to be able to purge the rotor bearings, if necessary, i.e. protect by inert gas, wherein this inert gas is fed from outside for each bearing position, then passes these bearings via a bypass bore and is again suctioned off at this collection/buffer space.
  • 14. Synchronization gearing as mechanical fall-back safe-guard for the electronic synchronization of the motor pair/spindle rotors, for example in the event of a power failure according to an operation by generator
  • 15. Control unit CU as control and regulation unit with evaluation of the particular current measurement values and output on this basis of the regulating signals for the intelligent operation of the spindle compressor with links and data preferably stored in the CU memory as well as continuously learning dependencies between the respective measurement values received and the gap values according to a previous simulation, verification and continuous experiences, the control unit is connected to FU-CU (24) as well as, on the side of the user, to process control technology for its application system as well as industrial control system in the sense of “Industry 4.0”
  • 16. FU control unit, designated as “FU-CU”, for the two frequency converters FU.2 (2.4) and FU.3 (3.4), wherein FU-CU exchanges the data regarding the operation of the spindle compressor directly with the control unit (15).
  • 17. Intermediate support on the two-toothed spindle rotor (2) with minimal thickness (below 2.5 g/cm³) yet high flexural rigidity, preferably as fibre composite material, e.g. CFRP vacuum-compatible
  • 18. Distance/spacer disks, preferably configured as so-called “peeling disks”, for the individual fixation of each spindle rotor in the longitudinal direction of the rotor axis for the selective adjustment of the gap value as Δ2.1 value on the 2t rotor (2) and as Δ3.1 value on the 3t rotor (3)
  • 19. Pressure-reducing organ in the closed internal circuit, for example via the utilization of the geodetic difference in altitude in a water tower, i.e. use of gravity for pressure reduction
  • 20. Thermal insulation for the intermediate water jacket (5)
  • 21. Angled rest for the control balls
  • 22. Centrifugal disk (preferably with rough/course surface) for feeding and fine distribution of the refrigerant fluid flow into the suction area (11) of the compressor according to K4 as injection cooling
  • 23. Feed of the refrigerant fluid flow K4 to the centrifugal disk (22), wherein this feed can be configured in accordance with one's wishes as:
  • 23.1 Feed of the refrigerant fluid flow K4 via a cantilever (24) of the bearing supports (25) 23.2 Feed of the refrigerant fluid flow K4 via the central bore in each rotor support shaft (2.1 and 3.1)
  • 24. Cantilever of the bearing support, preferably 3 for each bearing support (25), fluidically advantageous
  • 25. Bearing support for receiving the rotor bearings (4), simultaneously face-side chamber delimitation on the outlet side
  • 26. Regulating organ for the separation of the refrigerant fluid flows (9) managed by the control unit (15)
  • 27. End outlet openings with control edge (27.S) for each rotor after the control balls (10) as of a value regarding the inner volume ratio iV predetermined in accordance with the specific application
  • 28. Vacuum-compatible condenser housing vessel on the housing ø-dividing plate (1.P), attached in a horizontal position and sealed; in the case of a vertical operation, the condenser housing vessel is erect
  • 29. Vacuum-compatile evaporator housing vessel preferably attached on the housing ø-dividing plate (1.P) with insulation (29.i, illustrated as cross hatching or as an intermediate vacuum space) for ensuring the efficient Q_ent thermal balance; in the case of vertical operation, the evaporator housing vessel is located at the bottom
  • 30. Shielding-gas feed (so-called “purging”) for the protection of the bearings (4) as well as, if appropriate, of the drive units so that said areas to be protected constantly have a slightly (a few mbar of differential pressure, e.g. 3 to 10 mbar, are sufficient) higher pressure than the immediate R718-water-vapour surroundings; as inert/shielding gas, normal air will be sufficient in most applications, although, for example, nitrogen can also be selected
  • 31. Shielding-gas discharge generated via a separate vacuum backing pump, which preferably also ensures the vacuum system pressure. With this shielding-gas discharge system, a certain R718-water-vapour partial flow is also suctioned off via the working-space shaft seal so that this R718 loss must be re-fed to the system as water. As a seal for the reduction of this R718 water-vapour-loss partial flow, the flow resistance of the working-space shaft seal should be increased, for example via a brush seal
  • 32. Bearing-bypass bore for avoiding a gas flow through the mount (4)
  • 33. Feed of the refrigerant fluid flow for the cooling of the rotor interior for each spindle rotor as:
  • 33.1 Feed pipe in the central bore of each spindle-rotor support shaft
  • 33.2 Feeding of the refrigerant fluid flow K2 for the cooling of the interior of the 2t rotor via feed pipe (33.1)
  • 33.3 Feeding of the refrigerant fluid flow K3 for the cooling of the interior of the 3t rotor via feed pipe (33.1)
  • 34. Retainer bushing in central support-shaft bore to stop the refrigerant fluid from flowing off
  • 35. Symbol for the condenser according to the disclosure in accordance with FIG. 1
  • 36. Heat exchanger
  • 36.ü Heat exchanger on the cooling water side
  • 36.a Heat exchanger on the chilled water side
  • 37. Circumferential overflow groove in the evaporator for supplied evaporation water with drain (37.a) in the event of a mode of operation according to FIG. 13, wherein this draining water then becomes refrigerant fluid flow (9)
  • 38. Heat exchanger (e.g. as pipe system) for heat transfer Q_ent lying on the inner side in the evaporator (7)
  • 39. Spindle-rotor rotation unit for the three-toothed spindle-rotor system
  • 40. Spindle-rotor rotation unit for the two-toothed spindle-rotor system
  • 41. (Complete) compressor machine as open twin-shaft rotary displacement machine
  • 42. (Complete) R718 displacement compressor system