STEAM COMPRESSOR COMPRISING A DRY POSITIVE-DISPLACEMENT UNIT AS A SPINDLE COMPRESSOR

Abstract

The invention relates to a spindle compressor designed as a twin-shaft rotary displacement machine for delivering and compressing flow media, particularly steam. It comprises a pair of spindle rotors in a compressor housing (1) comprising an inlet collecting space (11) and an outlet collecting space (12). The centre distance of the pair of spindle rotors is at least 10% longer on the inlet-side end than on the outlet-side end. Each of the two spindle rotors (2, 3) is driven by an electric motor (18, 19), and an electronic synchronisation controls the electric motors (18, 19) such that the spindle rotors (2, 3) rotate in a contact-free manner.

Description

Cyclic processes are preferably described on the basis of Carnot's theorem, with heat output and heat absorption as well as a compressor as drive for the circulation medium in the gaseous phase. Cyclic processes are used very frequently and have become indispensable in our daily lives. These processes include clockwise and anticlockwise Carnot processes, with desired/targeted heat absorption to fulfil a cooling task (in the refrigeration and air conditioning field) or with desired/targeted heat output to fulfil a heating task (keyword “heat pump”) with heat exchangers for heat absorption and heat output. For the movement of the circulation medium, a drive in the form of a compressor for the circulation medium in its gaseous phase is generally required. Firstly, the circulation medium and its specific properties is critical. There are various artificial circulation media (generally chemically produced, such as HFCs and HFOs) and natural circulation media (such as ammonia, propane, propylene, isobutane, ethane.

Water, however, is irrefutably ideal as a circulation medium because of its general availability and the fact that it is completely non-toxic, can be safely used at low pressures in the form of steam, meets even the most stringent of guidelines and safety regulations, is resource-friendly, environmentally friendly, low maintenance, efficient and practically without any risk potential (incombustible, non-explosive, uncritical).

The challenge lies with the compressor, because in the working pressure range of a few mbar not only are enormous flow-rate volumes required, but also very high pressure conditions. This results in incredibly difficult compression conditions, in particular due to high temperatures, especially since the isentropic exponent for steam in this pressure range is quite high at about 1.327; by way of comparison, modern refrigerants lie in the region of just over 1.1 with correspondingly moderate temperature increases in the compressor.

The task of steam compression is nowadays performed by turbo compressors, but, in order for them to cope with the high pressure conditions, the compression must be performed in several stages with simultaneous intermediate cooling. Their fundamental characteristic weakness as turbomachines is that they allow only moderately satisfactory temperature and pressure conditions. If there were a more efficient compressor solution here, steam would be a significant advancement as a circulation medium because of its enormous advantages.

The object of the present invention is to provide the compression of (preferably) steam in the known working field and pressure range, which is generally referred to as a rough vacuum, by a positive-displacement machine which handles the desired pressure differences and the large p/p pressure conditions with the typically steep characteristic curve for a positive-displacement machine (i.e. pressure values over volume flow with the corresponding working points), wherein this machine must be completely dry running (no operating fluid) and should have a total efficacy for the entire system that is better for the entire field of application as compared to current turbo compressors, so that the user requirements in the field of refrigeration and in heat pumps as well as other (Carnot) cyclic processes are better met, especially in terms of a greater pressure range.

This object of compressing steam at pressures below atmospheric pressure (preferably between 6 mbar and 300 mbar, i.e. in the classic rough vacuum region) in a power range of less than 1 kW to well over 100 kW as refrigeration cycle power for refrigeration technology (i.e. industrial refrigeration, commercial refrigeration and building air conditioning) or as heat pump cycle power [the required compressor power is lower in accordance with the “COP” ^{(as example)} value] is achieved in the form of a 2-shaft positive-displacement machine according to the spindle compressor principle with a gas inlet chamber (11) and a gas outlet chamber (12), wherein the centre distance between the spindle rotors on the gas inlet side (11) is greater than on the gas outlet side (12) and thus results in a crossing angle alpha, which is preferably between 3 and 25 angular degrees, in such a way that the features described below are provided:

The features according to the invention are:

• 1) electronic synchronisation, since each spindle rotor (2 and 3) is driven by its own drive motor (18 and 19), each drive motor having its own FU (22 and 23), each with its own measuring system (20 and 21) for detecting the rotary angle position, and an FU control unit (24), which ensures that these drive motors (18 and 19) via their own frequency converter (22 and 23) are driven at a corresponding speed, such that the spindle rotor pair (i.e. 2 and 3) can work without contacting one another. The cooling fluid supply (9.2 and 9.3) to the cylindrical evaporator cooling bore (6) of each rotor is then provided additionally through the hollow shaft of the relevant drive, the bearings (10) then preferably being formed as grease-lubricated-for life hybrid bearings or all-ceramic bearings (or even as magnetic bearings).
• 2) cylindrical evaporator cooling bore (6) as a “rotating cylinder evaporator” for automatic cooling self-balancing, since the water to be evaporated, as spindle rotor cooling fluid under the pressure p₀* and the temperature t₀* [these values with certain technical deviations, such as pressure losses, temperature increase due to unavoidable heat transfers], is diverted from the circuit of FIG. 2 and, in the cylindrical evaporator cooling bore, by the rotary centrifugal forces, inevitably always goes during operation exactly to where it is currently most urgently required for the current working point, wherein the cylindrical evaporator cooling bore (preferably) has the following features in accordance with the following explanation:
  • A “rotating cylinder evaporator”, which is the design of the internal cooling of the spindle rotors in accordance with the invention, has the conceivably most favourable heat transfer properties for the problem addressed by the present invention, because the best-possible heat transfer is consistently achieved as a result of the centrifugal forces, since the heavy liquid parts in the rotating cylindrical evaporator cooling bore constantly displace the lighter gas components from the heat transfer surfaces in order to evaporate again immediately, so that the next liquid parts for the heat transfer thus reach the rotor material for the desired heat transfer, and additionally at the same time, still in the rotor longitudinal axis direction due to identical cylindrical radii values, the liquid parts to be evaporated are always displaced by centrifugation to the location where, due to the greatest evaporation, there is also the greatest need for heat dissipation, because in the rotor longitudinal axis direction different power distributions are provided for each operating point, so that, at the known high evaporation enthalpy differences with low (see values in FIG. 9) cooling fluid supply, the most efficient possible heat dissipation during the compression is achieved, so that in accordance with FIG. 8 the compressor line from to is advantageously steep and clearly is better for the compressor than an isentropic profile.
  • The following features apply for the cylindrical evaporator cooling bore:
  • a) Cylindrical evaporator cooling bore (6) of radius R_C along the length L_C with the spindle rotor positive-displacement profile length L_R, said cylindrical evaporator cooling bore preferably beginning between positions E and S in the inlet region and preferably going beyond the outlet end at L, so that the values for L_R and L_C are comparable (approximately equal). The cylindrical evaporator cooling bore (6) is designed as an “inner structure” preferably via cooling fluid guide grooves (16), cooling fluid distributor overflow grooves (17) and support points (7).
  • b) The cylindrical evaporator cooling bore (6) should be as cylindrical as possible (i.e. deviations well below 1%), wherein for example manufacturing tolerances in the R_C values are preferably set such that deviations tend to lead to larger R_C values in the direction of the outlet (i.e. in the region of position L).
  • c) Spindle rotor made of an aluminium alloy is rotatable conjointly with the “inner structure” already manufactured, wherein to form this “inner structure, the cylindrical evaporator cooling bore (6), preferably formed by cooling fluid guide grooves (16) with radius R_C and comprising multiple support points (7), is preferably pressed on to the supporting steel shaft at these support points, for example by component temperature difference, by joining the warmer aluminium rotor body to the cooler steel shaft and then making this a fixed connection with temperature equalisation, wherein only then is the gas-conveying “external thread” (31) produced, so that the wall thicknesses w can be minimised in order to improve the heat conduction through shorter paths in the event of dissipation of the compression heat.
    • The groove bottom of the cooling fluid guide grooves (16) is preferably designed such that the groove base surfaces are formed with inclination angles ψ(z), which, depending on the z-position in the rotor longitudinal axis direction, which is usually referred to as the z-axis, is preferably in the range
    • 170°≤ψ(z)≤180° depending on the position z in the rotor longitudinal axis direction, so that the distribution of the cooling fluid (9) with smaller amounts of cooling fluid (because there is always only so much cooling fluid supplied that the total energy balance at the particular working point gives the highest efficiency) along the rotor axis is improved by the smaller filling cross-sections depending on the current amount of cooling fluid, appropriate for the operating point. The cooling fluid guide grooves are designed here similarly to a thread, preferably with the greatest possible pitch, for example as in the case of the gas-conveying external thread (31), in order to thus perform the task of minimising the amplification of the residual unbalance resulting from the introduction of the cooling liquid (9.2 and 9.3) into each rotor (because all liquid collects in the rotating system at the greatest possible distance from the current rotation point and thus amplifies the residual unbalance), which is very poorly fulfilled for example with a pitch of zero of the cooling fluid guide grooves.
    • This effect of the residual unbalance amplification is used in accordance with the invention simultaneously to minimise the amount of cooling fluid supplied (9.2 and 9.3) per rotor, since vibration sensors (as used for example for bearing monitoring) display this residual unbalance amplification by an excessively large amount of cooling fluid in the corresponding rotor system, wherein, thanks to different rotor speeds (the 2t rotor always rotates 1.5 times faster), it is precisely determined at which rotor the amount of cooling fluid is too high, so that the control unit (25) can make the adjustment that is correct in the present case (in the sense of the minimum required amount of cooling fluid) via the regulation members (38).
  • d) To compensate for deviations and to ensure the most reliable possible distribution of the water to be evaporated in the rotor longitudinal axis direction at the cylindrical evaporator cooling bore (6), there are additionally also undersized cooling fluid distributor overflow grooves (17) in the bottom of the cooling fluid guide grooves (16) of radius R_C, which undersized cooling fluid distributor overflow grooves are arranged at a distance from the rotor axis of rotation which is greater than the R_C value, but at the same time have such a small cross-section that the water contained therein goes beyond the cross-section of these cooling fluid distributor overflow grooves (17) and wets the bottom of the cooling fluid guide grooves (16) of radius R_C.
    • The embodiment of the cylindrical evaporator cooling bore (6) is of course described here merely by way of example with support points (7) and cooling fluid guide grooves (16) of radius R_C and with cooling fluid distributor overflow grooves (17). Of course, other embodiments are also conceivable here.
  • e) Addition of cooling fluid (9) in particular to the rotors always limited to the minimum amount, possibly even only sporadically and in pulses, both to avoid critical unbalances and to minimise the amount of diverted cooling fluid flow (9) in the sense of maximising the overall efficiency, because this cooling fluid flow (9) lacks the actual circulation medium (28) in the evaporator (35) in the event of the heat absorption. The cylindrical evaporator cooling bore (6) in each spindle rotor thus receives only so much water (with a tolerance of for example+1%, which is conventional in this field) as is currently needed for evaporation at the particular working point.
  • f) This minimisation of the cooling fluid flow amount (9) is achieved for example by measurement via known and simple vibration sensors (for example for rolling bearing monitoring) for determining the degree of filling in the particular cylindrical evaporator cooling bore (6) per spindle rotor (2 and 3), because an increased amount of water in the particular cylindrical evaporator cooling bore (6) will amplify the residual unbalance in the rotating system, and, thanks to different speeds of the spindle rotors (the 2-toothed spindle rotor rotates faster than the 3-toothed spindle rotor by a factor of 1.5), as unbalance excitation can be associated with the rotation system of the 2-toothed or 3-toothed spindle rotor, so that the cooling fluid amount (9.2 and 9.3) is adjusted in accordance with the minimum amount. Thus, only as much water as is currently needed for evaporation in the current operating point is supplied.
    • Of course, other approaches for minimising the cooling fluid flow amount (9) can also be used.
• 3) Steam outlet (14) from the cylindrical evaporator cooling bore (6) for each spindle rotor, characterised in that each cylindrical evaporator cooling bore (6) is formed with the radius R_C2 or R_C3 and the steam outlet (14) is realised via transverse bores, which are preferably arranged balanced in relation to each other, after a step over a radius R_D2 or R_D3, wherein R_D2 and R_D3 at the particular spindle rotor is slightly smaller (i.e. a few mm, for example 2 to 5 mm) than the corresponding R_C2 or R_C3 value of the corresponding cylindrical evaporator cooling bore (6).
• 4) Cooling fluid injection (33) in the working space for selectively influencing the conveyed gas temperatures in the working space, i.e. the space between the inlet collection chamber (11) and outlet collection chamber (12).
• 5) With regard to the heat dissipation for the working space components, i.e. the pair of spindle rotors according to (2 and 3) and the compressor housing (1), said heat dissipation being so significant for the dry-running machine, two stages are to be distinguished:
  • A) Basic stage with component heat dissipation: The heat dissipation for the working space components can be used as a basis for safeguarding and ensuring at all times that play reduction (which generally leads to the failure of the compressor, or a “crash”) between the working space components is reliably avoided at all operating points:
    • This indispensable requirement is already achieved with small amounts of cooling fluid (9) by for example reducing the heat dissipation for the compressor housing (1), i.e. throttling the corresponding cooling fluid flow (9.1) with minimal cooling fluid flow amounts (9.1), so that the thermal expansion of the working space components does not jeopardise the play situation.
    • At the same time, in the case of this basic stage for component heat dissipation, it must be ensured that the play values (i.e. the distance values between the working space components) remain within a certain range, i.e. since the minimum play values during operation are about 0.03 to 0.09 mm (depending on the overall size, in large machines with >150 mm centre distance the values lie above 0.05 mm), the basic stage for component heat dissipation during operation should be configured so that not only is the aforementioned play reduction reliably avoided (as indispensable mandatory requirement, said minimum play values receiving a safety margin of about 20% to 50%), but also the play values for other operating points, due to the different thermal expansion behaviour of the components, compared to these lower play values are greater by a factor of 2 to at most a factor of 3, which is to be ensured by this basic stage for component heat dissipation during operation and is now attainable for the first time with a dry-running machine via the control unit (25) (previously only feasible with wet rotors).
  • B) FCT stage with component heat dissipation: (FCT stands for Final Compression Temperature, i.e. the temperature of the conveyed medium at the end of compression and usually the highest gas temperature, the FCT usually being determined in the outlet chamber (12).) The power demand during the compression of a volume (and that's exactly what happens in the present compressor provided in the form of a “positive-displacement machine”) is generally reduced, thereby improving the efficiency (efficacy) of the compression, if the temperature increase in this volume during the compression process can be minimised. The necessary heat dissipation during compression is known to depend also on the temperature difference between the gas in this volume and the surrounding heat-dissipating surfaces of the working space components, and also in addition the heat transfer coefficients (known in the case of steam to be quite high values) and the heat conduction (which is why an aluminium alloy is preferably used as material for the spindle rotors). Thus, the cooler the surfaces of the working space components can be kept via the cooling flow, the better is the heat dissipation during compression and the lower is the temperature increase of the conveyed gas in the conveyed and compressed working chamber volumes, hence the compressor working line becomes increasingly steeper—shown by way of example in accordance with FIG. 8 between the points and .
    • This generally achieves a reduction in the power demand for the compression and thus an improved (higher) efficiency.
• 6) Depending on the application-specific requirements and according to the corresponding parameter design (i.e. crossing angle, rotor length, inlet and outlet centre distance, head and root radii values per end-face section, gradient and number of stages as well as design of the “inner structure” and the spindle rotor pair cross-section) for the compressor design according to the invention, the cooling fluid flow (9) for heat dissipation for the working space components can be described by the following two approaches:
  • Diverted cooling fluid partial flow: (as shown by way of example in FIG. 2 as cooling fluid flow (9)) As shown in FIG. 2 by way of example, the cooling fluid flow (9) is diverted from the actual circuit as a partial flow, which is considered a preferred solution because it enables the maximum heat dissipation with the cylindrical evaporator cooling bore (6) during compression. The only disadvantage is the fact that this diverted cooling fluid flow (9) is removed from the main flow and thus is missing when the core task is performed in the field of refrigeration technology, i.e. the heat absorption in the evaporator (35). In heat pumps, when the heat output in the condenser (36) represents the core task, this diverted cooling fluid partial flow is not missing from the circulation medium (34).
    • Thus, the following principle applies: If, in a manner specific to the application, the advantage by reducing the compression temperatures in the mentioned FCT stage during component heat dissipation is greater than the disadvantage due to the reduced amount of cooling fluid (28) through the evaporator (35), then the diverted cooling fluid flow (9) should be realised by the cylindrical evaporator cooling bore (6) as shown by way of example in FIG. 2, wherein the amount of diverted cooling fluid flow (9) must be adapted in a targeted and controlled manner to the particular requirement profile in the sense that, in each situation and regulated by control unit (25), only such an amount is diverted as cooling fluid flow (9) that the compressor efficiency improvement by the heat dissipation brings more advantages in respect of the overall energy consideration than the previously described disadvantage, with associated additional effort by the diverted cooling fluid flow. If this approach is no longer achievable for some applications, then the “separate cooling water flow” described below applies.
    • Separate cooling water flow: (as shown by way of example in FIG. 6.d) If the advantage by lowering the compression temperatures in the mentioned FCT stage during component heat dissipation for the particular application is less than the disadvantage caused by the reduced cooling fluid amount (28) through the evaporator (35), then a separate cooling water flow as shown according to FIG. 6 with the internal rotor cooling described in PCT/EP2016/077063 should be realised, whereby it is ensured that play reduction between the working space components is avoided definitively and independently of the circulation medium.
    • The benefit that the separate cooling water flow for avoiding the play reduction somewhat lowers the compression temperatures quasi incidentally is indeed included. Naturally, the available cooling water temperatures are critical, and thus it is not possible to provide a generally valid guideline, and therefore a decision must be made separately for each specific application. Thus (in simple terms) the available cooling water temperatures will be different in a hot environment (countries near the Equator) than in cold regions at any given time of the year (Siberia in winter).
    • Delayed evaporation:
    • If evaporation of the cooling fluid (9.2 or 9.3) should not occur in the cooling fluid guide grooves (16) due to the enormous acceleration values, then it is further proposed in accordance with the invention that this cooling liquid (heated in the meantime by absorption of the compression heat) is drawn off by pitot tube (as described for example in DE 10 2013 009 040.7 or in 10 2015 108 790.1), has a higher pressure than p_C because of the high kinetic energy, and consequently at a point after the compression process, for example in the outlet collection chamber (12), is fed back to the circuit, where this liquid is then evaporated and can absorb heat again task-specifically, the amount of cooling fluid then being adjusted so that the overall efficacy is improved.
  • In any case, the correct cooling fluid amount (in terms of efficiency and unbalance minimisation) for the particular operating/working point of the control unit (25) is regulated, wherein the corresponding data are stored in this control unit (for example in accordance with appropriate process simulation). “Trial-and-error” is also used as a self-learning process, wherein the system itself tries out variations and uses the reactions (i.e. energy demand and net output) to itself determine the setting with which the highest efficiency is achieved for the current working point. This approach can also be referred to as an “action” approach. Therefore, it must be decided on a case-by-case basis which of these approaches best solves the application-specific task.
• 7) Adaptation of the inner volume ratio to the current operating point in accordance with the invention by additional partial outlet openings (15), which preferably open in a spring-loaded manner and allow a partial gas flow to escape from the particular working chamber into the outlet collection chamber (12) when, in this working chamber approaching the outlet, the pressure is greater than the pressure in the outlet collection chamber (12), so that in the working chamber a harmful over-compression (adversely affecting the efficiency) is avoided.
  • The inner volume ratio (i.e. the quotient of the working chamber volumes between inlet and outlet), as “iV value”, must be adapted in the best possible manner to the current operating point for the most efficient (i.e. energy-saving) compression, in order to avoid harmful over- or under-compression. During operation, the iV value can be adapted in accordance with the invention by means of additional partial outlet openings (15), but must first be determined via the spindle rotor pair design. The iV value is influenced fundamentally by the following 3 variables:
    • Centre distance between the rotor axes of rotation (variable due to crossing angle alpha between the rotor axes of rotation and larger at the gas inlet (11) than at the gas outlet (12).
    • Ratio of the rotor head radii to centre distance in the end section as μ(z) value, using the equations below at each point z in the rotor longitudinal axis direction, wherein in addition the root angle γ_F2 is selected purposefully to maximise the nominal pumping capacity, the exact procedure also being described in detail below.
    • Gradient in the rotor longitudinal axis direction (also determines the number of stages at the same time, i.e. the number of closed working chambers between inlet and outlet), wherein the rotor length is known to be as long as possible up to the critical bending speed. When determining the mentioned rotor pair parameters, the “inner volume ratio” as “iV value” (i.e. quotient of the ^[larger] inlet to the ^[smaller] outlet working chamber volumes) should thus be configured in accordance with the isentropic exponent of the conveyed medium, the compression process, in particular in respect of the heat dissipation during the working chamber volume change (i.e. the compression), and the desired compression ratio (i.e. the quotient of outlet pressure to inlet pressure).
    • By way of example, this relationship is presented with reference to the values mentioned in FIG. 7:
    • The conveyed gas (steam) is to be compressed, for example, from 7.0 mbar to 95.9 mbar, resulting in a compression ratio of:

95.5 divided by 7.0=13.7.

• • • Only in the case of isothermal compression (i.e. no temperature change during the compression) would an inner volume ratio of 13.7 be implemented here as well.
    • Due to the increase in temperature in the working chamber during the “polytropic” compression, in accordance with the invention, starting from the present isentropic exponent for steam in this range of about 1.327, an iV value of about 10 will be provided by the intensive heat dissipation during the compression via the cylindrical evaporator cooling bore (6), in order to avoid both over- and under-compression.
    • This iV value as a change in the working chamber volumes which results from the multiplication of the relevant spindle rotor pair cross-sectional areas and the extent in the rotor longitudinal axis direction (generally ascertained via the profile gradient), is now technically realised in accordance with the invention by:
  • a) varying the spindle rotor pair cross-sectional area in each end-face section (shown by way of example in simplified form in FIG. 3 as a planar sectional view) in the rotor longitudinal axis direction, wherein the inlet-side rotor pair cross-section is larger than the outlet side rotor pair cross-section. This cross-sectional change at the spindle rotor pair is now achieved by:
    • changing the centre distance via the crossing angle alpha of the two rotor axes
    • changing the profile tooth height via the mentioned μ(z) value at each z position
  •  This change in the cross-sectional areas at the rotor pair due to the change in centre distance and μ(z) value results in a “iv.aμ value”) (cf. FIG. 9), wherein the working chamber extent must be considered. In this case, it must be ensured that a cylindrical evaporator cooling bore (6), wherein each spindle rotor has its own R_C values, is provided alongside minimum wall thicknesses w in the supporting root main body (32) whilst simultaneously taking into account the different critical bending speeds, as presented by way of example in FIG. 9.
  • b) changing the profile gradient (generally referred to as m) in the rotor longitudinal axis direction: By changing the profile gradient, a “iV.m value” (cf. FIG. 9 as an example) is created, which is usually significantly (more than a factor of 3) greater than the “iV.aμ value”, wherein the number of stages (i.e. the number of complete working chambers between inlet and outlet) over the rotor length L_R still permissible from a bend-critical viewpoint whilst observing the “mesh limit” (what tooth gap depth still can be produced relative to the tooth gap width) must be taken into account.
    • Of course, these two changes act simultaneously and multiplicatively in the rotor longitudinal axis direction in order to arrive at the desired overall iV value, in this example=10, which is shown by way of example in FIG. 9.
    • It is known that the higher the total iV value is, the more intense is the heat dissipation during compression, wherein a reduction in the compression temperatures generally leads to an improvement in the compressor efficiency (i.e. increase in efficacy).
    • Now, if the working point deviates from these mentioned pressure values, the additional partial outlet openings (15) ensure the ideal adaptation to the current working points and thus an efficient compression process at any time.
• 8) Each spindle rotor (i.e. the aluminium part, which sits non-rotatably on the steel shaft) consists of 3 regions:
  • a) external gas-conveying thread (31) The external gas-conveying thread (31) is preferably made only after the connection, for conjoint rotation, to the steel shaft, in order to minimise the size of the root main-body wall thickness w.
  • b) root main-body wall thickness w (32) to minimise resistance to heat dissipation and to maximise heat dissipation accordingly.
  • c) “inner structure” consisting of cylindrical evaporator cooling bore (6) with support points (7) and lateral supports which are to be sealed off on the outlet side=for example by O-ring) and inlet-side steam outlet (14) for the cooling fluid evaporated in the cylindrical evaporator cooling bore (6) from the cooling fluid flow (9) per working space component.
• 9) By way of example, 4 positions in the rotor longitudinal axis direction for describing the spindle rotor design according to the invention are mentioned (of course, there may also be more or fewer positions, nevertheless the spindle rotor design according to the invention can be well described, wherein in FIG. 1 and FIG. 4 as well as FIG. 5 the following is true for the following positions from the gas inlet (11) to the gas outlet (12):
  • In this case, the following definition applies to each position in the rotor longitudinal axis direction (usually designated as z). “μ(z) value” per spindle rotor in the profile design at each z position for the gas-conveying external thread (31) of each spindle rotor:

$\begin{array}{c}{R}_{K2}\left(z\right)={\mu }_2\left(z\right)·a\left(z\right)\end{array}\text{or:}\begin{array}{c}{R}_{K3}\left(z\right)={\mu }_3\left(z\right)·a\left(z\right)\end{array}$

• • In addition, the root angle γ_F2 is purposefully selected by making it greater than 90° when μ₂>0.6, wherein the head cylinder width b_K2(z) does not fall below a selected limit value, for example 5 mm.
  • a) Position E:
    • on the rotor-pair end inlet side with the greatest distance between the spindle rotor axes of rotation as a_E value
    • in accordance with the invention with cylindrical flattened portion (27) on the inlet side of the 2-toothed spindle rotor (2) over the radius R_KE2 in order to expand the maximum/highest rotor head speed to a larger spindle rotor area, wherein preferably a radii-like transition, shown in FIG. 2 by way of example as “R·tan”, allows the uniform transitions.
  • b) Position S: (can also be represented as a range over several z-values)
    • with the largest μ-value preferably such that the inlet working chamber receives the largest possible volume without violating the stated boundary conditions (i.e. cylindrical evaporator cooling bore, wall thicknesses at the supporting root main body (32), blowhole freedom, critical bending speed etc.), wherein the μ-value according to FIG. 3 and the equations given for each z position in the rotor longitudinal axis direction are purposefully realised, as shown by way of example in FIG. 9.
  • c) Position V: (can also be represented as a range over several z-values)
    • wall thickness adapted in accordance with the tooth height with reduction of the cross-sectional area in order to realise the internal compression with simultaneous good heat transfer properties over the supporting root main body (32).
  • Position L: (can also be represented as a range over a plurality of z-values) preferably as a cylindrical end, which is expediently designed to protrude beyond the end of the external thread into the outlet chamber, as shown by way of example in FIG. 1. As an overview table, the preferred specific values for these positions are shown by way of example in FIG. 9. The emphasis is exemplary, because both other positions and also other values can be realised. The parameters mentioned in this FIG. 9 merely show a meaningful embodiment illustrating the “spirit” of this invention. In this case, each position can certainly also be implemented as a z-range over several z-values in the rotor longitudinal axis direction and not just as a singular z-position.
• 10) The crossing angle alpha according to FIG. 5 between the two spindle rotor axes of rotation is provided in combination with the particular μ(z) value in the rotor longitudinal axis direction such that each rotor has a cylindrical evaporator cooling bore (6) with minimal (i.e. in respect of the material strength matching the tooth height in question) wall thicknesses w on the supporting root main body (32) (for example, according to the above position descriptions of E, S, V and L), while taking into account the (preferably) blowhole-free rotor profile design of the gas-conveying external thread (31) and bending-critical speed “appropriate to the specific spindle rotor” (°*°) in accordance with the following point regarding the critical bending speed and provision of the inner volume ratio according to the embodiment previously described.
  • °*° “appropriate to the specific spindle rotor” means that, in accordance with the speed differences between the two spindle rotors, the 2-toothed spindle rotor rotating 1.5 times more quickly has both a higher flexural rigidity and a relatively lower rotational mass, so that the critical bending speeds are reached by both spindle rotors equally.
• 11) critical bending speed ω_critical for the two spindle rotors via their parameter design (i.e. in terms of diameter=stiffness) such that

$\begin{array}{c}{\omega }_{\underset{2\text{-}\mathrm{rotor}}{\mathrm{critical}}}=1.5·{\omega }_{\underset{3\text{-}\mathrm{rotor}}{\mathrm{critical}}}\end{array}\text{with}\begin{array}{c}{\omega }_{\underset{\mathrm{generally}}{\mathrm{critical}}}=\sqrt{\frac{c}{m}}\end{array}$

•  critical bending speed generally as square root of rigidity (incl. bearings) over mass.
  • To achieve high speeds, each rotor system in accordance with the invention is embodied as a rotation unit (40), as shown by way of example in FIG. 6b, this being of crucial importance because the balance is provided for the complete rotation unit (40), whereby the balancing quality is improved.
  • This is because it is well known that even well-balanced individual parts, which are later assembled to form a rotation unit which can no longer be balanced separately as a unit (which is practically always the case with prior art 2-shaft positive-displacement machines), in their sum then give a poorer balance quality than the separately balanced and henceforth unchanged rotation unit, as shown by way of example in accordance with the invention in FIG. 6b.
• 12) Play adjustment between the spindle rotor and the compressor housing via separator plates (26) by first introducing each spindle rotor, during assembly, individually into the compressor housing (1) until the spindle rotor heads contact the housing bore and then pulling them out again and fixing them via the separator plates (26), so that exactly the desired head gap value between the rotor head and housing is given, as shown in FIG. 6c by way of example as Δ2.1.
• 13) The following rules must be observed for the bearings:
  • Since the bearings are only a single element in contact and therefore subject to wear, the bearings must be designed with particular care. Therefore, the following rules for the bearings must be observed:
  • In the case of steam, the bearing forces (both axial and radial) are very low and the main load is caused by the high speed, which is why in bearing technology what is known as the n·d_m factor is used as speed characteristic, i.e. the product of mean bearing diameter [in mm] multiplied by the speed [in rpm=1/min], wherein the machine tool construction under the keyword “spindle-bearing” here provides precise design recommendations. If this speed characteristic exceeds one million mm/min, special emphasis must be placed on speed resistance and lubrication. The rotor speed results from the maximum permissible rotor head speed below supersonic for the conveyed medium in the working area. The limit value for steam in the pressure range is specified at about 400 m/sec, which is why in accordance with FIG. 9 a value with sufficient safety margin is selected in the table with 350 m/sec, by way of example. According to the invention, the 2-toothed spindle rotor (2) is also flattened cylindrically in the inlet region so as not to hit the speed limit too early in this region, because in the outlet direction the rotor head velocity drops rapidly due to smaller diameter values (see FIG. 9, the table values).
  • For the present invention, the bearings are for example/preferably to be constructed as hybrid spindle bearings (e.g. of type XCB70) sealed on both sides with appropriately adapted lifetime lubricant and distanced appropriately from the conveyed medium via the working space shaft seals, wherein these working space shaft seals, in addition to separation and defence means (see ima catalogue from machine tool construction for spindle seals), also have neutral collection/buffer chambers (13) as protection as well as the imperative(!) avoidance of any gas flow through the bearings, which invariably need a safe bypass, i.e. a gas-permeable bypass (channels, holes) with minimal flow resistances. Instead of the aforementioned hybrid spindle bearings, of course, all-ceramic bearings are also feasible, as well as magnetic bearings if appropriate, and even water bearings.
• 14) The evaporator cooling for the working space components can be represented as a horizontal line with the pressure p₀* at t₀* according to FIG. 8, as shown by way of example in FIG. 2: For differentiation, the * are used, because this pressure can definitely and specifically differ from the pressure p₀ at t₀ in the evaporator (35), if advantageous according to the application-specific process simulation. It is also possible to perform the evaporator cooling for the working space components via a separate refrigeration cycle.
• 15) Instead of the rotor pairing with 2-toothed spindle rotor (2) and 3-toothed spindle rotor (3) as “Tribivari”, other rotor pairings are also conceivable (although probably less efficient), such as: rotor pairing as “SynchroVari” according to DE 10 2016 004 048.3 as well as the classic 2: 2-cycloid rotor pairing (but with blowhole)
• 16) The control unit (25) satisfies the particular application-specific requirements in that the control unit (25) manages and intelligently regulates, controls and monitors the entire system. All relevant data are stored in the control unit (25) and are collected and evaluated.
• 17) The displacement machine according to the invention, hereinafter referred to simply as “Tribivari”, is designed as an intelligent system, which is solved by the features and properties described below, where the abbreviation “ES” stands for the “electronic motor pair spindle rotor synchronisation” according to the invention. This novel intelligence can be presented by the following special tools:
  • (self) diagnostic tools
  • regulatory tools
•  The following somewhat more detailed explanations are intended to facilitate comprehensibility, although there may inevitably be some repetitions and “embellishments” under the consideration from different perspectives, with slightly differing terms (due to the different perspectives, which certainly improves the comprehension).

A compressor works in principle between the following two limits:

• • efficient compression (minimising internal leakage, suitable π_i value, effective heat dissipation, etc.) as a soft limit
  • Avoidance of gap reduction (crash) as a hard limit
  • Challenge for these limits: (This applies especially for dry runners)
  • A) individual for each individual machine (manufacturing tolerances, assembly differences etc.)
  • B) change during the runtime (deposit formation, dirtying, wear, etc.)
  • C) dependent on the particular operating point (in particular pressure range, volume flow, etc.)
  • D) vary with changing ambient conditions (hotter, colder, dirtier etc.)
  • Summary:
  • The more precisely the individual limits of each compressor in its particular situation are known and useable(!) during its service life, the better this system will be.
  • What can Tribivari do better than today's compressors?: Today's compressors (in particular as dry-running machines) are designed to survive the worst-case scenario, i.e.: At the other working points they are inferior because of higher leakage.
  • Tribivari, by means of “PartCool”, manages at all times(!) the thermal situation of all components for the compressor efficacy and can thus adapt to “all” conditions with ongoing self-diagnosis and Δ-compensation! (Δ stands generally for deviations and differences)
  • “PartCool”=cooling water flow guided intelligently by CU for each component, plus harmonisation of inner compr. ratio
  • The thermal situations of all components are finally known individually(!) and at all times and are adjustable by setting the relevant cooling fluid flows selectively by algorithm in the CU.
  • How does Tribivari provide its individual°*° intelligence?:
  • °*° individual=for every machine in every situation & environment at any time
  • check the k₀ speed** and continuously compare with stored value
  • measure flow resistance by Σpressure difference degradation as a function of time
  • inverse cooling to determine the crash security for ΔT as a temperature difference
  • measured value comparison with subsequent extrapolation
  • . . . etc. . . . .
  • The CU is used for the specific PartCool regulation thanks to an algorithm with targeted compensation of deviations and differences, which learns by comparison.
  • Tribivari knows at all times how far its own load can be driven in each case
  • to: a) safely avoid a crash (gap reduction)
  • and b) to intelligently maximise the compression efficacy for precisely this “situation”
  • incl. c) to learn independently by its own comparison(!):
    • What was good? What was bad?=>This leads to the relevant optimum
  • plus d) by extrapolation as a prognosis with corresponding notification “upwardly” (outside).
    i.e.: Tribivari helps itself by fixing itself practically single-handedly.

What is the new Tribivari CU “intelligence”? [CU=control unit]

Unlike the currently only “semi-mounted” (screws already worked beforehand) controllers, the intelligence of Tribivari is a conceptual part of this new compressor technology, since the entire operation is managed by the control unit individually (i.e. specifically to each Tribivari with its own tolerances and particular use conditions/deviations) under all conditions, including constant changes thereto, with independent self-diagnosis(!) and prognosis with ongoing adaptation to the process under various conditions (colder/hotter environment, poorer cooling, etc.). This is the new Tribivari intelligence.

Today's compressors can only adapt inadequately to the current process and changes thereto and to changing environmental conditions (e.g. hotter). Reasoning:

• • A) “oil-injected screws”=injected oil quantity (indispensable due to internal leakage, heat dissipation and lubrication) cannot be adjusted arbitrarily in respect of oil quantity and oil temperature.
  • B) “dry-compressors”=they do not manage all(!) of the thermal situation of their working space components and therefore have only one good working point (minimum gap) for crash avoidance (gap reduction) and otherwise work “unhappily” at extreme speeds.
  • C) In addition, none of these machines can adapt their internal compression ratio (i.e. between over-compression and under-compression) to the current operating point (cf. refrigerant compressor effort for housing slide)

Tribivari is fundamentally superior here in that it fulfils 3 features simultaneously:

• • Tribivari=dry PLUS eta PLUS μC What is key is the PLUS of these features.
  • This is because Tribivari manages the thermal situation of all working space components, as follows:
  • any time=the CU permanently monitors the compressor and always regulates the cooling fluid flows via what is known as PartCool (as described)
  • complete=both with regard to process and environmental conditions as well as all appropriate working space components
  • flexible=different and changing conditions during the process and in the surrounding environment are tolerated
  • comprehensive=via cooling fluid mass flow and cooling fluid temperature suitable for any current situation and not only for one working point, but always optimal for the entire working area
  • synchronous=the working space components are always managed synchronously (=always in step, no divergence)
  • efficient=always with appropriate heat dissipation (and not according to the motto: “the more the better”, but in each case as appropriate=intelligent!), the best polytropic exponent and no over-/under-compression for desired pressure volume flow
  • intelligent=with own learning(!) algorithm with self-diagnosis and prognosis, even forward-looking
  • Specifically with Tribivari:
• a) monitoring and management of the thermal situation of all working space components
• b) so that thanks to ACTUAL gap differences, the inner leakage (=entropy) can also be handled via the gap dimensions
• c) always adjust the inner compression ratio appropriately with additional partial outlet openings
• d) optimise the magnitude of heat dissipation to the polytropic exponent of the compression
• e) and thus maximise the efficacy
• f) adapt or monitor the temperature level in an application-specific manner if necessary
• g) to give the desired conveyed gas amount and the setpoint operating pressure by rotor speed and cooling fluid adjustment.

What, among other things, forms part of the “Tribivari_CU-intelligence”?:

• • (better than the currently only “semi-mounted” FU intelligence in the case of screws)
• 1) Play values Δ: =gap distances between the working space components:
  • Δ2.1=2-t rotor from the housing
  • Δ3.1=3-t rotor from the housing
  • Δ3.2=rotors from each other
  • Δ as function ƒ(z) possible in rotor longitudinal axis direction
  • Δxy denotes integral for all gap distances

Purpose:

• • reliable avoidance of gap reduction (=crash),
  • knowledge of the size of the leakage gap (for highest efficiency per working point)
  • . . . that's not possible today
• 1.1) Detecting the actual individual gap values (in particular with regard to manufacturing tolerances and assembly differences) in each AirEnd assembly via separator plates on the fixed bearing exactly set for each(!) rotor according to:
  • a) for Δ2/3.1 by contact+withdrawal
    • ( . . . as an assembly-must per rotor, alas integral over l, not in operation . . . )
  • b) for Δ2/3.1 by inverse cooling
    • ( . . . also in operation after the k₀ speed measurement, PartCool division)
  • c) for all Δxy by k₀ measurement
    • ( . . . clear, with PartCool adaptation, also in operation plus inverse cooling)
  • d) for Δ2/3.1 with hot rotors
    • ( . . . inaccurate temp. level/fluctuations/duration)
  • e) for all Δxy by spying
    • ( . . . barely accessible)
  • f) for Δ3.2 by electronic synchronisation ( . . . note: Δφ of the emergency synchro-wheels is detected first)
• 1.2) Detecting the changes of these gap values during operation according to A) to D) as ongoing conclusion regarding comparison of measured values and/or k₀ speed and/or flow resistance and rotation angle Δ with electronic synchronisation . . . =all with interpolation and extrapolation
• 2) Self-diagnosis:
  • Evaluation by algorithm in the CU based on the individual gap values according to 1) with determination of need for action incl. tendency detection (prognosis) with intelligent analysis plus vibration sensors (especially for bearing monitoring)
• 3) Adaptation to the process, especially in the case of process changes:
  • (=much more than just today's speed adjustment) Adaptation to the particular process and its process changes by evaluating the differences in the algorithm and leading to actions=for example PartCool adjustment
• 4) Adaptation to the environment, especially in the case of environmental changes:
  • Adaptation to different and changing environmental conditions
• 5) Efficacy Optimisation:
  • always lowest possible energy consumption through optimal cooling regulation
  • not just for a single working point (as before), but for the entire area
• 6) Temperature control:
  • Compliance with desired limit temperatures (esp. sensitive process gases)
• 7) Compression adjustment:
  • Change in internal compression ratio via additional partial outlet openings (15) to avoid under- and over-compression
• 8) Purity of the conveyed medium:
  • Adaptation of the size of the secondary gas flow to the neutral chamber per working area shaft seal
• 9) BASIS:
  • Simulation algorithm stored in the CU, fed by the individual gap values and the current situation and correspondingly adapted reactions, based on maps that are constantly being expanded, interpolated and compared(!), with constant learning.
• 10) Electronic synchronisation:
  • each spindle rotor with its own (synchronous) motor plus encoders,
  • Cooling fluid fed to the spindle rotor cooling thread through the hollow motor shaft (with single clutch)

Tribivari Helps Itself by Fixing Itself Practically Single-Handedly, i.e.:

Tribivari is intelligent insofar as Tribivari with the mentioned (self-)diagnostic tools in the form of “self-diagnosis” firstly recognises, itself, if Tribivari changes due to wear, abrasion, dirtying and/or deposit formation, and can then adjust its operating behaviour on that basis via the described control tools, specifically this means that, for example at each operating point, as required by the user's process operating point in the particular situation,

a) the most appropriate gap values are set via PartCool or PartCool&Control,
b) each optimal inner compression ratio is set via post-inlet and/or pre-outlet,
c) and the most suitable rotor speed is set.

Based on the individual start state of this Tribivari system stored in the control unit, the current state (due to wear, abrasion and/or dirtying, deposit formation possibly changed) of Tribivari is taken into account in the algorithm of the control unit, as are also the current ambient conditions (hotter, colder, dirtied heat exchangers, etc.) and the currently desired operating requirements (i.e. in terms of volume flow, pressure level, but also allowable power consumption in the sense of avoiding expensive power peaks, etc.).

EXAMPLE 1

Tribivari uses its own (self-)diagnostic tools, i.e. by means of k₀ speed measurement and/or ΣΔρ measurement and/or algorithm-measured value comparison and/or Δφ rotor pair check and/or inverse cooling, etc., incl. any (evaluation) combination of these tools, to determine that the gap values have decreased in the outlet area, for example by deposit formation/dirtying. Tribivari can determine this via the algorithm in its own control unit, where individual guideline values (stored during the assembly of this Tribivari) are available for the different measured values and are stored with the respective links, relationships and interpretations, which are then compared with the incoming measured values. The control unit then adjusts the regulating tools as a regulation unit for this Tribivari system, for example in that PartCool reduces the cooling for the compressor housing via the outlet-side cooling fluid flow (9.1a) and/or intensifies the two cooling fluid flows (9.2 and 9.3) to the spindle rotors. If these diagnostic results (here as an example: gap values at the outlet are reduced) were not known, there would be a risk that Tribivari would continue to cool the working space components and thus increase the risk of gap reduction (=crash). Thanks to this approach according to the invention via the mentioned (self-)diagnostic tools and regulating tools, these limits are now known for each operating point and use conditions, and Tribivari can not only be operated safely, but also in the optimum (in the sense of minimal energy requirement) range. At least by means of inverse cooling, it can even be determined individually for each spindle rotor which gap value has decreased, namely Δ21 or Δ3.1, in order accordingly to increase the relevant cooling fluid flow 9.2 or 9.3 according to the value tables present in the CU (for example previously calculated by FEM simulations).

EXAMPLE 2

Tribivari determines via its own (self-)diagnostic tools that the gap values have increased in the inlet area, for example due to abrasion/wear, noticeable by way of poorer compression behaviour. To compensate for (to “rescue”) this situation, for example, the cooling fluid flow 9.1b at the compressor housing inlet area must be increased.

Based on “PartCool” as a self-diagnosis by means of:

• • k₀ speed measurement and/or ΣΔρ measurement combined with
  • inverse cooling (at least as a safety check against a crash situation)

Content and Purpose:

Measurement of the compressibility of a compressor at zero flow rate (i.e. only “counteracting” the internal leakage and not expelling any conveyed medium at the outlet) for different rotor speeds within the scope of the Tribivari CU intelligence for the purpose of:

a) determining the actual achieved individual compression quality level at the end of assembly as a control and Okay approval (i.e. within the desired tolerance) stored in the system's own CU as a base output reference for continuous comparison during operation for the purpose of detecting a tendency and for prognosis displayed by extrapolation.
b) Self-diagnosis during operation to detect changes (for example caused by wear, abrasion, dirtying, deposit formation, operational changes, for example in the process and/or in the environment, etc.)
c) Preferably, the k₀ speed measurement is possibly also combined with ΣΔp measurement with the inverse cooling as an ongoing operational check for reliable crash avoidance by means of extrapolation.

Procedure for k₀ Speed Measurement:

If the inlet pressure is known, the outlet (over)pressure achieved is measured at the closed outlet for different rotor speeds and, thanks to PartCool, at defined(!) thermal situations**°° of the relevant (i.e. in particular the working space) compressor components (and the resultant individual gap conditions), and the quotient of outlet-to-inlet pressure gives the desired k₀ speed value for this rotor speed, and thus as a value table or as a functional representation:

y-axis=k₀ value as quotient p_a/p_i

x-axis=rotor speed n_R

• **°° Because, thanks to the CU intelligence by means of PartCool&Control, the thermal situation, and, via the thermal expansions of all working space components, the gap values can be regulated and controlled in a targeted manner, in the case of the k₀ speed measurement, the individually defined component temperatures are used to determine the particular compression quality level, and the comparison with the base output reference values as well as further measurements during operation reveals not only the current state but also the changes:
  • i.e. self-diagnosis as well as prognosis and tendency. In addition, the sufficient safety margin for crash avoidance is determined by means of the inverse cooling, that is to say for the stated working limits:
    • both safe crash avoidance
  • and
    • also the most efficient compression possible

Inverse cooling=simulation of a “wrong” (inverse) component cooling with a component temperature difference, as no longer occurs later during operation (because constantly monitored by the CU in this sense as well)

Both the k₀ speed measurement and the inverse cooling are repeatedly used during operation to detect changes within the lifetime of this compressor.

Simplification:

The inverse cooling is also executable via an algorithm stored in the CU as extrapolation of several “harmless” (in the sense of readily available) hot-fluid temperatures (preferably from the warm fluid reservoir (33), for example).

First as an overview: (then explained separately)

The following (self-)diagnostic tools belong (by way of example) to the Tribivari intelligence:

1) contact+withdrawal+fixation
2) inverse cooling
3) k₀ speed measurement
4) ΣΔp measurement
5) algorithm-measured values comparison
6) Δφ rotor pair check
7) combination & evaluation
8) . . . etc. . . . further (inherent) diagnostic tools can also be added here)

And the following regulating tools belong to the Tribivari intelligence (by way of example):

A) “PartCool”, also called “PartCool&Control”
B) π_i adaptation
C) FU speed variation

D) “ActionStep-ReactionCheck”

E) combination & evaluation

In Tribivari spindle compressors according to the invention at least the temperatures mentioned in FIG. 1 are measured, not only from the cooling fluid but also from the components. This is very easy with the compressor housing and in the frame-fixed inlet and outlet area, because these are stationary (frame-mounted) components. In the case of the rotating spindle rotors, the relationships between cooling fluid temperatures and rotor temperature for the various load states are stored in the control unit (25), so that the “defined temperature conditions” described hereinafter for the entire Tribivari spindle compressor are always known sufficiently precisely in the CU (25) or can be converted via interpolations (known geometry and material properties) with the resultant individual gap conditions also widely known.

These “defined temperature conditions” are an ongoing prerequisite for the correct use of these tools, which is ensured with sufficient accuracy thanks to the extensive temperature measuring points. (preferably similarly simple sensors such as those common in today's automobile construction industry and in widespread use)

Since the temperature conditions are never exactly the same, there is installed in the CU an algorithm for conversion to a uniform comparable state, which will henceforth be referred to by the term ‘defined temperature conditions’.

Separate cooling fluid temperature ranges at the reservoir (10) facilitate the achievement of defined temperature conditions by removing cooling fluid selectively for the relevant component.

The following (self-)diagnostic tools belong to the Tribivari intelligence (by way of example):

• 1) “Contact+Withdrawal+Fixation”: (This only occurs at the time of assembly of the compressor stage) During assembly, each finished°*° spindle rotor is introduced individually into the compressor housing (1) until complete contact with its housing bore as so-called “zero-gap”, that is to say touching, is achieved, wherein it must be ensured that contact between the rotor and housing is as complete as possible (check as appropriate by means of touch paste and rotate slightly by hand to secure the rotor-housing contact), and therefore the housing is preferably upright and the spindle rotor is introduced from above. Because the (mean) inclination angle γ₂ or γ₃ between spindle rotor and housing bore is known, this rotor must now be pulled out again in the rotor longitudinal axis direction over a path distance Δz_path that can be directly calculated by trigonometry and fixed between the inlet cover (16 or 17) and the compressor housing (1) via adjustable distance/spacer plates (34 or 35) in order to satisfy the desired (mean) gap value Δ2.1 or Δ3.1 between spindle rotor (2 or 3) and the compressor housing (1), wherein according to FIG. 6.a the following is true:
  • for the 2t rotor:

$\begin{array}{c}\Delta z_{\mathrm{Path}2}=\frac{\Delta 2.1}{\sin \left\{{\gamma }_2\right\}}\end{array}$

• • and for the 3t rotor:

$\begin{array}{c}\Delta z_{\mathrm{Path}3}=\frac{\Delta 3.1}{\sin \left\{{\gamma }_3\right\}}\end{array}$

• • It must be ensured that the components (i.e. the spindle rotor in question and housing) have approximately the same component temperature, which also must be logged or must be taken into account when entering the data into the CU memory (also to be entered into the CU). Advantageously, the gap size Δ2.1 and Δ3.1 can thus be set and logged in a targeted manner, which hitherto has not been possible. In this case, a constant inclination angle γ₂ or γ₃ is advantageous, however different inclination angles are also possible in accordance with expansion laws in the rotor longitudinal axis direction (i.e. according to simulation of the compression process and heat dissipation of the working space components), and therefore a mean inclination angle can be applied, or the inclination angle that, according to the simulation of the compression process and heat dissipation of the working space components, primarily defines the gap dimension Δ2.1 and Δ3.1.
• *°*° finished spindle rotors: in the form of a rotation unit (40) with the corresponding inlet cover (16 or 17) fully assembled, wherein in particular the fixed bearing (10) is important for this process.
• 2) Inverse Cooling:

In the “inverse cooling”, the gap dimensions between the working space components are measured and checked in that, at minimum (or even zero=standstill) speed of the spindle compressor,

• • a liquid (for example water) with steadily increasing fluid temperature is conducted in a controlled manner through the cooling fluid regions of each spindle rotor (2 or 3) preferably in some sections via the transverse bores (29)
  • and or
  • a liquid with steadily decreasing fluid temperature is conducted in a controlled manner through the various cooling fluid areas of the compressor housing (1) preferably in some sections,
    where the post-rotatability of the spindle rotors is checked constantly, for example manually, at the time of assembly or for self-diagnosis according to the invention in later operating pauses by electronic motor pair-spindle rotor synchronisation. Because of the different thermal expansions of these working space components, the rotatability of the spindle rotors will be terminated at a specific temperature level for this spindle compressor machine, and the specific ACTUAL cold play values for these spindle compressors are known on the basis of the known material properties and the known geometry conditions and are stored in the CU (25) for this spindle compressor.

Instead of using the first touching as “post-rotatability limit”, however, at least one ΔT_BT previously defined for this spindle compressor machine size should be established as the setpoint component temperature difference value and should ensure, via the simple (slow) rotatability monitoring, that there is no contact (touching) of the working space components. During later operation of this spindle compressor, the control unit (25) then knows how to adjust the particular cooling fluid flows of the working space components such that this ΔT_BT value is not exceeded, and therefore the crash can always be reliably avoided. This targeted regulation of the individual screw compressor components is also referred to below as “temperature control”.

In order to also be able to determine, in addition to the simple post-rotatability of the spindle rotors, the actual situation of the gap values and how this is affecting the compression behaviour of the Tribivari system, in accordance with the invention the inverse cooling is additionally performed by ΔT_BT value examination also linked with measures as described under “Combination & Evaluation”.

The ΔT_BT values, amongst other things in order to safeguard the crash avoidance, are thus considered to still be reliable, and component temperature differences checked multiple times by inverse cooling are continuously observed and preserved by the CU during operation in that the ΔT_BT values in question are not exceeded.

In particular, for higher compressor powers (for example over 75 kW power), it is useful to use the inverse cooling partially selectively in the rotor longitudinal axis direction by (as shown in FIG. 1) controlling the temperature selectively of individual areas both on the spindle rotor side and on the housing side using fluid, so that it is clearly recognisable how different the gap values in the rotor longitudinal axis direction are.

With these values, the “PartCool” can then be adjusted in a regulated manner during later operation in such a way that the gap values in each area are optimal: Optimal means that on the one hand a crash (i.e. gap reduction) is safely avoided, which is now finally possible thanks to the knowledge of the respective ΔT_BT values, and on the other hand the internal gap leakage can be monitored via the gap values managed by PartCool in accordance with the present simulation of the compressor process in such a way that the efficacy is maximised for precisely the current compression process.

For inverse cooling, the following case distinctions are expedient:

• a) Assembly-related inverse cooling:
  • Here, at the time of assembly, the original start state is recorded specifically for each spindle compressor after contact+withdrawal+fixation, with the actual ΔT_BT values as mounting ΔT_BT values, and is stored in its control unit (25). In addition, linked measures are carried out as described under “Combination & Evaluation” and these individual measured values are stored in the CU for this Tribivari compressor. This process forms the reference for possible changes (due to wear, abrasion, dirtying, deposit formation, etc.) during later operation.
• b) Use-related inverse cooling
  • In the case of use-related inverse cooling, the assembly ΔT_BT values (or similar in order to be able to determine the assembly ΔT_BT values on the basis of a stored algorithm in the CU*°°*) are preferably repeated and linked with Combination & Evaluation to identify the current state of this Tribivari system.
  • The use-related inverse cooling is performed preferably during breaks in operation, wherein the fluid with higher temperature for the fluid flow areas of each spindle rotor comes from a warm fluid reservoir (33). For this warm fluid reservoir (33), a cooling fluid partial flow is either diverted uncooled during operation and “warm parked” there or is selectively heated there by electric heating generated there.
  • Of crucial importance here is the comparison of the currently determined values with the previous values, in order to be able to operate the Tribivari system optimally (i.e. avoiding a crash and at the same time with best efficacy) and also identify trends and enable forecasts.
  • *°°* If this process of drawing conclusions is adequately backed up by sufficient experience, the use of warm fluid can be spared later via extrapolation and interpretation. Nevertheless, it sometimes helps to produce the defined temperature conditions.
• 3) k₀ speed measurement:
  • In the case of the k₀ speed measurement, the instantaneous compressibility of this spindle compressor machine is integrally checked, wherein in particular the changes in the algorithm of the control unit are evaluated in the sense of adaptation of the PartCool and recognition of a trend.
  • Measurement of compressibility of a compressor at zero flow rate (i.e. only “counteracting” the internal leakage and not expelling conveyed medium at the outlet) for different rotor speeds within the scope of the Tribivari CU intelligence for the purpose of:
    • Determining the actually achieved individual compression quality level at the end of the assembly as a control and with OK approval (i.e. within the desired tolerance values) in the inherent CU as base output reference stored for constant comparison during operation for the purpose of identifying a trend and for prognosis by extrapolation.
    • Self-diagnosis during operation to determine changes (for example caused by wear, abrasion, dirtying, deposit formation, operational changes, for example in the process and/or in the environment, etc.)
    • Preferably, the k₀ speed measurement is possibly also combined with total pressure difference measurement combined with inverse cooling as an ongoing operational check for safe crash avoidance by means of extrapolation.

Procedure for k₀ Speed Measurement:

If the inlet pressure is known at the closed outlet for different rotor speeds and thanks to PartCool under “defined temperature conditions” (Explanation=see above), the outlet (over)pressure) that has been reached is measured and the quotient of outlet pressure to inlet pressure gives the sought k₀ speed value for this rotor speed, and thus in the form of a value table or as a function representation, for example according to:

y-axis=k₀ value as quotient p_a/p_i

x-axis=rotor speed n_R

• **°° Because, thanks to CU intelligence, the thermal situation and, via the thermal expansion of all working space components, also the gap values can be selectively regulated and controlled by PartCool, the compression quality level is determined via the individually defined component temperatures for the k₀ speed measurement, and not only the current state, but also the changes are identifiable via the comparison with the base output reference values and other measurements during operation: i.e. self-diagnosis and prognosis and trend (by extrapolation).
  • In addition, the sufficient safety margin for crash avoidance is determined via the inverse cooling—for the stated working limits:
  • that is to say, both reliable crash avoidance and the most efficient compression possible.
  • The k₀ speed measurement as well as the inverse cooling are repeatedly used during operation to detect changes in the service life of this compressor.
• 4) integral pressure difference measurement, abbreviated as “ΣΔp measurement”:
  • During the ΣΔp measurement, the current flow resistance of the Tribivari compressor stage is measured by setting a selected overpressure at “defined temperature conditions” with open inlet and closed outlet in the outlet collection chamber (12) and by measuring the reduction of the pressure in the outlet collection chamber (12) for a selected period of time (for example 3 minutes) with very slowly rotating (for example less than 10 revolutions per minute) spindle rotors. This individual ΣΔp measurement takes place for the first time at the end of the assembly of each AirEnd spindle compressor and is stored in the CU as a “Base reference”. During the course of operational use, this ΣΔp measurement is repeated in the breaks in accordance with a selected rhythm controlled by the CU and is compared both to the base reference and to all follow-up measurements. From this, a prognosis and trend can be deduced by extrapolation.
• 5) Algorithm Measured Value Comparison:
  • During operation, there are many measured values, regulatory actions and reactions in the CU (25) as well as various evaluations. Based on the previously performed simulation calculations as well as (FEM) model calculations of the relevant compressor components, a steadily growing database is created, which is continued with the constantly incoming data. In the CU (25), these data are now continuously compared with one another and interpolated using an algorithm, so that they are also mapped (“modelled”) and stored for instances of use that are not exactly the same as those currently occurring (for example, higher conveyed gas inlet temperatures); the CU (25) delivers the appropriate output signals (32.e).
  • Initially, due to the even smaller amount of data, the comparison and interpolation will initially still be rough and will be provided fuzzily with an increased level of uncertainty; however, as the individual (rather specific inherent) database grows steadily in the CU, this fuzziness will becomes less and the machine will get increasingly better and smarter.
• 6) Δφ rotor pair check: with Δφ as the angle of rotation between the spindle rotors
  • In the case of the Δφ rotor pair check, the gap situation Δ3.2 between the spindle rotors (2 and 3) is checked individually for each Tribivari system via the electronic motor pair-spindle rotor synchronisation by means of the rotary angle sensors (20 and 21) for each shaft strand measuring the exact rotary angle play and comparing it both with the base reference attained at the time of assembly as well as the follow-up measured values, and evaluating this in the sense of identifying trends and providing a prognosis.
  • Procedure:
  • In the electronic motor pair-spindle rotor synchronisation, one motor string (i.e. each spindle rotor with carrier shaft fixedly connected to its drive motor rotor shaft) is then electrically blocked (thus fixed) and the other motor string then checks the remaining rotary angle Δφ as “remaining rotary angle play” and stores this value. This measurement is repeated a number of times for the entire spindle rotor pairing, and the maximum and minimum values are stored and compared again (with base reference and follow-up measured values), in order to (if the values are correct, i.e. lie within the stored tolerance ranges) set the mean value as setpoint specification for operation by electronic motor pair-spindle rotor synchronisation. With the electronic motor pair-spindle rotor synchronisation, it is possible to determine for the spindle rotor pair via
    • reduced rotary angle play values a deposit formation or dirtying
    • increased rotary angle play values an abrasive wear or surface abrasion and wear
  • with appropriate feedback by CU, for example, to higher-level maintenance and service stations.
• 7) Combination & Evaluation:

The (self-)diagnostic tools mentioned are not only to be used and evaluated individually, but in particular also in combination. For example, the inverse cooling does not have to be driven until the first contact of the working space components as a check of the post-rotatability limit (also because of the risk of surface damage), in that the post-rotatability is ensured with a ΔT_BT of the inverse cooling defined previously in the CU (i.e. a clearly defined temperature level of the working space components) on the one hand, and on the other hand in that a ΣΔp measurement and/or k₀ speed measurement are/is performed, wherein the values then determined by these methods are compared with the basic reference and comparison values that are appropriate for this inverse cooling and are stored in the CU.

And the following regulating tools belong to the Tribivari intelligence (by way of example):

• A) “PartCool”, also known as “PartCool&Control”:
  • The most important regulating tool is the individual control and regulation of the cooling fluid flows for each component over the relevant amount (mass flow) of the cooling fluid flow as well as over the cooling fluid temperatures. This is not a “bullish” control, but a control or regulation implemented by the system response having a direct influence on the mentioned PartCool parameters, hence the extended name of “PartCool&Control”. Virtually all changes in the Tribivari system as well as in the process and in the environment can be compensated by PartCool&Control, because thanks to the data stored in the CU for the particular work process (even if “only” as extrapolation or interpolation of directly available data) as well as the expansion behaviour and resulting gap values with corresponding internal gap leakage values etc., the corresponding compression behaviour of the Tribivari system can be optimally adapted in each case.
• B) π_i adaptation:
  • Each work process experiences different conditions (for example in respect of pressure and temperature values, volume flow, ambient conditions, etc.), so that adjustments to the compression process are desirable for the desired minimum compression energy requirement. These adjustments also include the “inner compression ratio” as the inner π_i value of the compressor machine, which at first purely geometrically describes the ratio of the inlet chamber volume to the outlet chamber volume. The actual compression ratios (in particular the temperatures and heat dissipation during compression) result in the known over- and under-compression, which should firstly be minimised as far as possible. The control unit of the Tribivari system according to the invention can now adjust the internal compression ratio of the current situation in an ideal manner at any time via the regulation of partial gas flows via additional partial outlet openings (15): This regulating tool in operation is referred to as “π_i adjustment”.
• C) FU speed variation:
  • With this classic and long-known procedure, the spindle rotor speed is adapted to the particular conditions by FU ^{(=frequency converter)} in particular with regard to the currently desired conveyed medium volume flow: as is known, almost proportional to the rotor speed.
• D) “ActionStep-ReactionCheck”:
  • Guided by the control unit at selected intervals (for example a few minutes), small changes are made continuously, for example in the cooling fluid flow to a working space component such as the compressor housing and/or in the π_i adjustment, etc. It is important that the ΔT_BT values stored in the CU for safe crash avoidance are always observed (=important!). Based on the constantly incoming measured values (in particular temperatures), it can now be determined in the algorithm of the CU whether this change has resulted in an improvement or deterioration in the current compression process, in particular determinable via the energy consumption (i.e. engine torques and engine speeds and/or also only the motor current consumption).
  • Thus, “ActionStep-ReactionCheck” is a constant self-learning and iterative method, which can be regarded both as a (self-)diagnostic tool and as a regulatory tool, because the system responses also reveal conclusions about the current state of the Tribivari system.
• E) Combination & Evaluation:
  • The aforementioned regulating tools are not only to be used and evaluated individually, but in particular also in combination. Thus, for example, PartCool&Control and π_i adaptation via the CU's own algorithm are always carried out in a coordinated manner, preferably evaluated and performed in combination by ActionStep-ReactionCheck. The storing of the results in the CU's own database constantly increases the knowledge of this Tribivari system and thus forms part of the Tribivari intelligence.

In the case of the Tribivari intelligence, the evaluation of measures carried out is an essential precondition for the above-mentioned regulations.

This evaluation is carried out in accordance with the following features:

with the abbreviation “ES” as “Electronic motor pair-spindle rotor synchronisation”
for example in respect of “ActionStep-ReactionCheck”:
Improvements are noted when, at a working point at an existing pressure p_B, the power requirement (which is even known at ES for each rotor) is reduced or minimal at a known speed, wherein the gap leakage and entropy balance are assessed in the algorithm of the CU via the temperature feedback (32.e) thanks to simulations and ongoing learning (writing of the “experiences” of this machine), so that the compr. efficacy can be specified. This is henceforth referred to as the objective of an efficient compression process for the current situation.

A volume flow measurement for the conveyed medium is generally too time-consuming, but would provide a nice facilitation if it were carried out or were available. Instead of improvements, of course, deteriorations in the compression behaviour also can be noted and are evaluated by the control unit, in order then to be able to initiate appropriate regulation measures (in particular by PartCool&Control etc.).

Advantageously, the cooling fluid flows are regulated by the CU in an application-specific manner for the particular situation in accordance with the algorithm stored in the CU and flexibly based on current experience by seeking the relevant optimum, wherein in particular the sufficient heat dissipation via a conventional external heat exchanger with advantageous temperature differences is taken into account.

This is the new Tribivari intelligence according to the invention and is not easily possible in the prior art.

The Tribivari system knows practically at any time with sufficient accuracy, how its individual status is currently (=how it is exactly, for example in respect of dirtying, deposit formation, state of wear, load capacity, temperature level, current gap values and compressibility etc.) in order to be able to use this knowledge to optimally perform the particular work process in the current situation(!) (optimally in the sense of most efficient compression in the current situation(!)).

In addition, thanks to the trend and forecast analyses mentioned, the CU will forward its status in good time to higher-level service and maintenance positions in order to permanently ensure the upkeep, care, maintenance and service as well as the availability of this system.

The Tribivari system is designed to be self-learning by continuously updating the analysis data individually for each CU under the particular process conditions and continuously optimising them further using ActionStep-ReactionCheck and storing them in the CU's own database.

Command Values for CU Intelligence: (!)=important (-)=less important

• 1) (!) Cooling fluid flow to 2t rotor
  • (by speed of its own cooling fluid delivery pump or the regulating member)
• 2) (!) Cooling fluid flow to the 3t rotor
  • (by speed of its own cooling fluid delivery pump or the regulating member)
• 3) (!) Cooling fluid flow to the housing=can be metered per section (in particular for larger machines, for example >75 kW)
• 4) (-) Cooling fluid flow to the side parts (actually only to the outlet side part)
• 5) (-) Cooling fluid flow to the lubricant (no longer with electronic synchronisation)
• 6) (!) Rotor speed (via FU=frequency converter:
  • possible without slip=synchro-motor)
• 7) (!) Additional partial outlet openings as variable partial gas flows

Measured Variables:

• a) almost all temperatures
  • in particular practically all temp. differences ΔT for each cooling fluid flow and at the conveyed gas plus lubricant temperature as well as the component temperatures (in particular on the housing and the side parts)
• b) rotor speed
• c) torque per rotor (with electronic synchronisation)
• d) each cooling fluid mass flow
  • (at least via the selectively regulated speed of the cooling fluid delivery pump/characteristic possibly precisely) results with ΔT the heat dissipation per working space component in each working point at any time
    Special features:
• 1) The actual gap values Δ2.1 and Δ3.1 and Δ3.2 are detected at the time of assembly of the compressor individually per machine, for example by “inverse cooling” or “Contact+Withdrawal” and stored in the CU, wherein these values during operation are then aligned with the algorithm in the CU to regulate the different cooling fluid flows for each working space component, such that on the one hand a crash (i.e. gap reduction) is reliably avoided (depending on the size of the machine, for example with about 15% safety reserve) and on the other hand the gap values do not exceed a specified maximum value (depending on the machine size, for example, about 1.5 times the cold gap values recorded from the time of assembly).
  • This is because the CU is always aware of the compressor state due to the management of the thermal situation and the thermal expansion behaviour of the working space components stored in the CU.
• 2) Self-diagnosis and prognosis
  • for determining, recording and evaluating the current state, in particular with electronic synchronisation to the speed variation
• 3) Process adjustment & ambient adjustment & temperature control & compression adjustment via additional partial outlet openings
• 4) k₀ speed measurement combined with inverse cooling for the purpose of targeted determination of the current (i.e. corresponding to the current state) individual PartCool with PartCool&Control

Definition of “working space”=space between inlet (11) and outlet (12) The working space is defined by the pair of spindle rotors (2 and 3) and the surrounding compressor housing (1) with the narrow (in the region of 0.1 mm and smaller) gap values Oxy of the various components. In the working space, the desired compression of the conveyed medium takes place via the working space components, i.e. spindle rotor pair (2 and 3) and compressor housing (1).

In addition, it is also true in accordance with the invention that the CU as a control unit (25) not only monitors, regulates and optimally manages the spindle compressor as described, but at the user location also not only communicates (for example Profibus system) with the entire system/plant controller via automation technology as industrial controller in the “process management technology”, but also actively participates therein, for example by managing/regulating the load management for the entire (at least in the case of this user) system, consisting of the individual compressor systems with their own CU (25) in each case, and therefore for example costly current peaks are avoided, wherein this then belongs to the term “Industry-4.0”. This also includes (preferably) at the same time (if the user agrees) also feedback to the supplier (or to the suppliers if there are more than one) regarding the current state of the compressor system with all individual systems including prognosis for further conduct with appropriate maintenance recommendation regarding the known diagnostic systems (for example vibration sensors, temperature profiles, etc.) with the corresponding evaluations (software). In addition, this also includes the constant and continuous adaptation to changed or changing process conditions, for example by deposit formation, dirtying, wear, etc., but also by external environmental conditions such as temperature level (for example warmer or colder environment), another desired pressure level, whereupon the intelligent CU system (25) responds by appropriate adjustment of the cooling water amounts, equalisation of the inner compression rate by means of additional partial outlet openings (15) etc., as well as all measures for self-diagnosis to determine the current state of this compressor in this application and prognosis over the further course with appropriate remedial action ranging from adjustment of the cooling fluid amounts up to a warning to the operator.

In the figures, instead of a subscript, merely a dot is inserted as index, so that for example R.F2 means R_F2 and thus here denotes the root radius on the 2-toothed spindle rotor, wherein:

	F	stands for	profile root
	K	stands for	profile head
	C	stands for	cooling
	WK	stands for	pitch circle
	2	stands for	the 2-tooth spindle rotor (2)
	3	stands for	the 3-toothed spindle rotor (3

FIG. 1 shows, by way of example, a 2-toothed spindle rotor (2) in longitudinal section with rotor geometry according to the invention and with cylindrical evaporator cooling bore (6) according to the invention and adapted positive-displacement profile root-base wall thickness w for the load-bearing root-base body (32) on the basis of the example of the 2t rotor with detail of the steam outlet (14) over a plurality of (balanced with the necessary cross-section Σ) transverse bores from the cylindrical evaporator cooling bore (6) with the radii values which are as follows:

R_w2<R_D2<R_C2

for the preferably blowhole-free profile pairing, the gas-conveying “external thread” (31) on the 2-toothed spindle rotor is located above the pitch circle line (37). As is known, the drive motor (18) consists of a motor rotor (mounted on the carrier shaft 4 for conjoint rotation) and a motor stator assembly with the electrical stator motor windings (shown by squared hatching),
optional: extraction to the vacuum pump (29) starts at the neutral chambers (13) of the working space shaft bushings, in order to protect the bearings from the conveyed medium as necessary

FIG. 2 shows, by way of example, a cooling circuit with diversion of to cooling fluid (9) from the circuit, with cooling fluid injection (33) into the compressor working space per working point, targeted adjustment of the inner compressor volume ratio as iV value by additional partial outlet openings (15), with steam outlet (14) per working space component, i.e. housing (1) and rotor pair (2 and 3), shown in the inlet space (11)

the expansion valve, which is also shown, in the case of steam as the circulation medium, is preferably replaced via the simple height difference with the use of gravity as a “hydrostatic pressure difference” (the present illustration would then have to be adapted to the direction of the force of gravity).

The control unit (25) receives and processes various signals regarding the current operating requirements, the entire circulation system and in particular also from the compressor according to the invention, in order in particular to adjust the compressor components for each working point via the regulation members (38), such that the requirements are met in the best possible way—only with the control unit (25) can the system work reliably and efficiently (in practice a “New Intelligence”).

• • Referring to PCT/EP2015/062376=similar, but now improved by said inventive features to meet the requirements of steam.

FIG. 3 shows, by way of example, a spindle rotor pair end-face section with an adaptation of the μ(z) values in the rotor longitudinal axis direction simplified as a projection in a common plane, because the rotor axes of rotation are at the angle alpha to each other and ought to be shown three-dimensionally, for the various positions E, S, V and L of FIG. 5

where the following is true for the μ(z) values:

$\begin{array}{c}{R}_{K2}\left(z\right)={\mu }_2\left(z\right)·a\left(z\right)\end{array}\text{and:}\begin{array}{c}{R}_{K3}\left(z\right)={\mu }_3\left(z\right)·a\left(z\right)\end{array}$

Adaptation of the μ(z) values for the rotor pairing according to the invention, preferably as 3:2 pairing to fulfil the following 3 core tasks:

• • maximising the nominal pumping capacity (based on the rotor pair cross-sectional area, achieve the greatest possible scoop area)
  • with blowhole-free rotor pairing (minimise internal leakage)
  • with optimum use of the critical bending speed at each spindle rotor, specific to their respective speeds

Design: For each of the 2t rotor and the 3t rotor with different cooling bore ø values R_C2 and R_C3 wherein the supporting steel shafts have not been shown for simplicity, and

different head strength distribution, since the root angle γ_F2>90°, so that the tooth cross-section of the 2-toothed spindle rotor (2) is slightly slimmer, without dropping below a minimum head width b_K2 below (for example 5 mm).

This happens in such a way that the critical bending speeds per rotor (i.e. for 2t and 3t) match, so that the following is achieved for the spindle rotor pair:

• • The rotor pairing is without a blowhole, and therefore the internal leakage is reduced.
  • Based on the illustrated rotor pair cross-section, this design achieves significantly more scoop area and thus an increased pumping capacity relative to the cross-section, which is sought for steam compression.
  • Accordingly, the 2-toothed spindle rotor has the larger cooling bore for heat dissipation during compression, so that the component heat balance is improved in respect of heat absorption and heat dissipation.
  • The 2t rotor has a speed 1.5 times higher than the 3t rotor and accordingly it is embodied in accordance with the invention in such a way that this 2-toothed spindle rotor achieves the more rigid shaft thanks to R_F2>R_F3 at reduced (by means of γ₀>90°) mass, which is favourable for the increase of the critical bending speed, because the 2-toothed spindle rotor also has to rotate faster and accordingly has to be designed in accordance with the invention with the higher critical bending speed limit.
  • Accordingly, the slower 3t rotor has a lower bending critical speed due to the lower bending stiffness, for which reason it also rotates slower.
  • According to the invention, the rotor pair is now designed in such a way that the critical bending speed at the 2t rotor is 1.5 times higher than the critical bending speed at the 3t rotor, wherein the following is sought:

$\begin{array}{c}{\omega }_{\underset{2\text{-}\mathrm{rotor}}{\mathrm{critical}}}=1.5·{\omega }_{\underset{3\text{-}\mathrm{rotor}}{\mathrm{critical}}}\end{array}\text{with}\begin{array}{c}{\omega }_{\underset{\mathrm{generally}}{\mathrm{critical}}}=\sqrt{\frac{c}{m}}\end{array}$

bending critical speed generally as the square root of stiffness over mass

FIG. 4 shows an example as shown in FIG. 1 but for the 3t rotor with external profile conveying thread area below the pitch circle line (37), displacement profile area=where there is arranged the outer conveying thread (31) with profile teeth and tooth gap areas, which form the various working chambers as a series connection between the inlet and outlet and below the pitch circle line (37) ensure the blow hole-free compression.

FIG. 5 shows by way of example: rotors from FIG. 1 and FIG. 3 paired to show the overall rotor geometry and indicating the crossing angle alpha and the spindle rotor pairing with the engagement lens area engaging centrally with one another

FIG. 6 shows, by way of example, a total of 4 CAD illustrations, showing:

• 6a) a compressor housing (1) formed as a “pot housing”:
  • i.e. outlet-side closed bottom side and internal processing of the working space from the open inlet side
• 6b) Rotation unit:
  • each spindle rotor with carrier shaft, bearing, drive motor and measurement system as a completely assembled and balanced unit (40), ready for mounting and henceforth unchanged, shown here only with the example of the 2-toothed spindle rotor, although the same applies for the 3-toothed spindle rotor, wherein the cylindrical flattened portion (27) at the 2t rotor inlet is not shown.
• 6c) assembly and play adjustment:
  • shown for both via separator plates (26) for the important rotor head play relative to the housing, by way of example as a detail for the head play Δ2.1 between 2-toothed spindle rotor head and housing. The final clearance adjustment between rotor heads and the housing is performed via separator plates (26), this being illustrated by way of example as Δ2.1 in FIG. 6c for the 2-toothed spindle rotor head.
• 6d) finished machine:
  • Both rotation units mounted in the pot housing plus frequency converter (22 and 23) per motor incl. FU control unit (24), which communicates with the control unit (25) for continuous data exchange, which FU control unit in turn is connected to the user process controller.
  • The motor windings of the two drive motors (18 and 19) are protected against the conveyed medium for example by vacuum-proof potting of the motor stator winding assemblies or also by gap pots between the motor stator and motor rotor, etc.

The rotor internal geometry according to FIGS. 1 to 4 with the cylindrical evaporator cooling bore has not been included in FIG. 6, since this embodiment, as described, instead of applying for the described evaporator component cooling via a cylindrical evaporator cooling bore (6) according to FIG. 2, now applies for the option with separate cooling water flow as cooling water operation according to the industrial property right PCT/EP2016/077063, wherein in this embodiment a cylindrical internal rotor cooling is not required, because the internal rotor cooling shown in FIG. 6 is suffice.

This FIG. 6 shows:

• • good and reliable balancing for the rotary units, in particular for the desired high speeds implemented in the case of steam to about 350 m/sec as max. rotor head speed.
  • easy installation as a modular system, since different rotor pair variants in the same housing geometry
  • targeted play adjustment via the separator plates (26) in order to be able to compensate for the particular tolerance situation (because all production parts have deviations/dimensional differences within certain tolerances) caused by unavoidable manufacturing tolerances “individually” (as precisely as possible for these various components).
  • electron. synchronisation via (18) and (19) as drive for each rotation unit
  • and with μC as a control unit for the intelligent cooling of the components (as described above)

FIG. 7 shows by way of example: operating/working points as a basis (Excel) for the prior art=for turbo, improvement by the present invention by the higher ΔT with heat dissipation for t_C

• • more ΔT is desired for heat release below t_C
    • this cannot be done by one of today's turbos (already working with 2 stages)
    • there must be a positive-displacement machine, which creates the p/p pressure ratio
    • at the same time the machine must imperatively be formed as an absolute/complete dry-running machine due to steam

FIG. 8 shows, by way of example, an illustration of the compression process in a pressure-enthalpy graph in the case of steam compression, showing the improvement due to the intensive evaporator heat dissipation during compression

• • prior art is shown as the dot-and-dash line (with labelling)
  • improvement according to the invention is shown as dashed line (with labelling) compressing from to

Purpose of the Presentation:

Prior art represented by turbo, which must work in two stages with intermediate cooling, as compared to the improvement of the invention, here referred to as “HydroCom” (abbreviated to HC)

Explanation of the Prior Art:

In order to isentropically compress^(Carnot) from 8 mbar (to =4° C.) to 48 mbar (t_c=32° C.), intermediate cooling is indispensable for a 2-stage turbo because already isentropically from 8 mbar to 48 mbar there would already be a temperature rise of from 4° C. to approx. 200° C., without intermediate cooling.

Improvement According to the Invention:

Because of the enormous p/p pressure conditions with high isentropic exponent, the best-possible heat dissipation during compression must be ensured, which would otherwise lead to a fatally high (in the sense of increased compressor power) rise in compression temperatures, and therefore in accordance with FIG. 8 compression is performed practically almost at the dew line (i.e. better than isentropically), wherein the rotor pair cooling effort for to somewhat worsens the overall efficiency in refrigeration technology due to the diverted cooling fluid flow (9.2 and 9.3).

Thus, HC fulfils a stronger requirement profile according to FIG. 7 in that, in accordance with the invention, improved HC works from 7 mbar=2° C. to 96 mbar=45° C., thanks to efficient heat dissipation during compression.

FIG. 9 shows, by way of example: an Excel design table with example values for the parameter values for the positions E, S, V and L, stated by way of example, in the rotor longitudinal axis direction for the spindle rotor pair with individual values per spindle rotor, the indicated power specifications being only quite rough and constituting provisional reference values. Of course, both the selection of the named positions and the selection of other parameter values for the particular application-specific requirement profile are imperative.

Therefore, it should again be emphasised at this juncture that this is merely an example, showing only one of many possible design options for the rotor pair design according to the invention for demonstration purposes only.

For some applications, it may be favourable that the cylindrical evaporator cooling bore (6) is designed in a multi-stepped cylindrical form, as “terraces” so to speak, with the overflow edge as shown by way of example in FIG. 1.

Where reference is made here generally to cooling fluid, what is meant here is the R718 cooling fluid known from the field of refrigeration, which is naturally compressed at the chosen negative pressure as steam in the positive-displacement machine according to the invention, or in liquid form as cooling fluid (9) for component cooling by evaporation.

Terms such as “substantially”, “preferably”, and “the like”, and also “possibly”, which are understood to be imprecise, are to be understood such that a deviation by ±5%, preferably ±2%, and in particular ±1% from the normal value is possible. The applicant reserves the right to combine any features and also sub-features from the claims and/or any features and also sub-features from a sentence in the description in any manner with other features, sub-features or partial features, even 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 reference signs, so that they are generally described only once.

Since the lowest temperatures in the case of steam are above 0° C., the combination with the refrigerant R744 as CO₂ (as a 2-stage solution, also known as a “cascade”) is advantageous for lower temperature values (for example for deep freezing).

The invention relates to steam compression for refrigeration, air conditioning and heat pump technology, both for clockwise and anticlockwise (Carnot) cyclic processes. In order to improve the efficacy and operating behaviour at the same time with a greater pressure range, the present invention proposes a dry 2-shaft positive-displacement machine as spindle compressor, the spindle rotors (2 and 3) of which have a rotor pair centre distance which on the inlet side (11) is at least 10% greater than on the outlet side (12), and being driven by electronic motor pair (18+19)-spindle rotor (2+3) synchronisation, and each spindle rotor being provided with internal cooling, wherein the crossing angle alpha between the two rotor axes of rotation is formed in combination with the corresponding μ(z) value in the rotor longitudinal axis direction in such a way that a preferably cylindrical evaporator cooling bore (6) with minimal wall thickness w at the supporting root-base body (32) is formed for each spindle rotor under simultaneous consideration of the (preferably) blowhole-free profiling of the gas-conveying external thread (31) and critical bending speed “appropriate for the specific spindle rotor” and implementation of the inner volume ratio as iV value, wherein the inner volume ratio is adjusted during operation via additional partial output openings (15) and the gas-conveying external thread (31) in the case of a 2-toothed spindle rotor (2) is preferably formed with a cylindrical flattened portion (27) in the inlet region.

LIST OF REFERENCE SIGNS

• 1. Compressor housing with outer cooling areas and inlet-side greater distance of the spindle rotor receiving holes than on the outlet side, these bore axes being preferably intersecting (i.e. with perpendicular distance zero) or also crossing (or skewed), with external cooling fins for a cooling fluid flow rate (9.1) managed by control unit (25), preferably with cooling fluid flow, for example according to (9.1a) and (9.1b), in some sections in the rotor longitudinal axis, wherein for larger rotor lengths (for example >500 mm) a plurality of cooling fluid flow-through sections are formed on the compressor housing, and the compressor housing preferably is embodied as a so-called pot housing according to FIG. 6a.
• 2. Spindle rotor, preferably with 2-toothed gas-conveying external thread (31), called a “2t rotor” for short, preferably made of an aluminium alloy with good thermal conductivity (preferably above 150 W/m/K), fixed for conjoint rotation via support points (7) on a steel shaft (4) and inside having a cylindrical evaporator cooling bore (6) with radius R_C2.
• 3. Spindle rotor, preferably with 3-toothed gas-conveying external thread (31), called a “3t rotor” for short, preferably made of an aluminium alloy with good thermal conductivity (preferably above 150 W/m/K), fixed for conjoint rotation via support points (7) on a steel shaft (5) and having inside a cylindrical evaporator cooling bore (6) with radius R_C3.
• 4. 2t-rotor carrier shaft, connected to the 2t rotor for conjoint rotation at radius R_W2 (preferably pressed on) with central cooling fluid supply bore (4.a), preferably integrally and at the same time also shaft for the 2t drive motor (18)
• 5. 3t rotor carrier shaft, connected to the 3t rotor for conjoint rotation at radius R_W3 (preferably pressed on) with central cooling fluid supply bore (5.a), preferably integrally and at the same time also shaft for the 3t drive motor (19)
• 6. Cylindrical evaporator cooling bore with radius R_C and length L_C for the corresponding spindle rotor, preferably with cooling fluid guide grooves (16), cooling fluid distributor overflow grooves (17) and support points (7)
• 7. Support points as a rotationally fixed contact between spindle rotors (2 and 3) and carrier shafts (4 and 5).
• 8. Synchronisation toothing for the spindle rotor pair, also rotating in the case of electronic synchronisation as fallback transmission for emergency situations, for example power failure, wherein the motors then automatically switch to generative operation and only at the end (own power generation is no longer enough) does the transmission prevent the spindle rotor contact.
  • As a fallback transmission, no lubricating oil is required, wherein this toothing is realised with increased overlap ratio (i.e. larger toothing bevel angle) so that the profile overlap can be reduced by decreasing the tooth heights for smaller sliding motions in the tooth engagement to reduce friction and hence wear, wherein the tooth flanks preferably still receive a dry-running coating as protection.
• 9. Cooling fluid flow for cooling the compressor working space components, i.e. rotor pair and housing, either diverted from the circulation medium (34) according to the example in FIG. 2 or as a separate cooling fluid flow shown in FIG. 6d generally, wherein, for example the following is true:
• 9.1 Cooling fluid flow to the compressor housing, for greater rotor lengths (for example >500 mm) divisible into:
  • 9.1a cooling fluid flow through a portion of the compressor housing (for example housing outlet side)
  • 9.1b cooling fluid flow through another portion of the compressor housing (for example central area)
• 9.2 cooling fluid flow to the 2t rotor
• 9.3 cooling fluid flow to the 3t rotor
• 10. Spindle rotor fixed bearing for receiving the gas pressure axial forces and for exact fixing of each spindle rotor in the longitudinal axis direction
• 11. Conveyed gas inlet collecting space for the conveyed medium with the gas pressure p₀ (for simplification, pressure losses in the lines are initially ignored)
• 12. Delivery gas outlet collecting space for the conveyed medium with the gas pressure p_C (for simplification, pressure losses in the lines are initially ignored)
• 13. Neutral collection/buffer space per working space shaft passage with reduced gas pressure with respect to the system pressure, preferably for example generated by negative pressure/vacuum pump.
• 14. Steam outlet via several transverse bores after a step with radius R_D2 or R_D3 per rotor
• 15. Additional partial outlet openings as diverted conveyed medium outlet partial gas flow with a regulating member (pressure difference valve) for adjusting the internal volume ratio
• 16. Cooling fluid guide grooves with the radius R_C per cylindrical evaporator cooling bore (6) with groove base surfaces at an angle of inclination ψ, which is preferably 170°≤ψ≤180°, and the cooling fluid guide grooves as a thread with the greatest possible pitch=as in (31)
• 17. Cooling fluid distributor overflow grooves (with undersized cross-section) preferably in the groove bottom of (16)
• 18a. 2t drive motor as a direct drive for the 2t rotor, preferably embodied as a synchronous motor
• 19. 3t drive motor as a direct drive for the 3t rotor, preferably embodied as a synchronous motor
• 20 Rotary encoder for measuring the exact rotary angular position of the motor 2t rotor carrier shaft (4)
• 21. Rotary encoder for measuring the exact rotary angular position of the motor 3t rotor carrier shaft (5)
• 22. Frequency converter, referred to as “FU.2”, for the 2t drive motor (18)
• 23. Frequency converter, referred to as “FU.3”, for the 3t drive motor (19)
• 24. FU control unit, designated as “FU-CU”, for both frequency converters FU.2 (22) and FU.3 (23), wherein the FU-CU directly exchanges the operating data with the control unit (25).
• 25. Control unit CU as a control and regulation unit with evaluation of the current measured values and output, based thereon, of the regulation signals for intelligent operation of the spindle compressor with links and data preferably stored in the CU memory as well as ever-learning dependencies between the incoming measured values and the gap values according to previous simulation, verification and ongoing experience, the control unit is connected to FU-CU (24) as well as the user side with the process control technology for its application system as well as factory control in the sense of “Industry 4.0”
• 26. Distance/spacer plates, preferably embodied as “separator plates” for individual fixing of the spindle rotor in the rotor longitudinal axis direction for targeted gap value adjustment as Δ2.1 value on the 2t rotor (2) or as Δ3.1 value on the 3t rotor (3)
• 27. Cylindrical flattened portion (as “cyl.” dimension specification in FIG. 2) on the 2-toothed spindle rotor (2) over the radius R_KE2 on its rotor inlet side
• 28. Circulation medium through the evaporator (35) for heat absorption (as a core task in refrigeration technology)
• 29. Vacuum pump for removal of foreign gases and for generation of the necessary negative pressure for the steam cycle, preferably sucking said gases into the neutral spaces (13) to protect the (rotor) bearings.
• 30. Water reservoir to compensate for water losses
• 31. Gas-conveying external thread with preferably blowhole-free profile rotor pairing to perform the compressor core task, namely to transport the gaseous conveyed medium from the inlet (11) to the outlet (12) and at the same time compress it
• 32. Supporting root-base body with wall thickness w at each spindle rotor (2 and 3)
• 33. Cooling fluid injection into the working space of the compressor
• 34. Circulating medium through the condenser (36) for heat output (as a core task in heat pumps), circulating medium here is steam (circulating through different states), but in principle also suitable for other circulation media, for both clockwise and anticlockwise Carnot processes
• 35. Evaporator for the circulating medium, in which a quantity of heat is absorbed.
• 36. Condenser for the circulating medium, in which a quantity of heat is output.
• 37. Pitch circle line (abbreviation: WK) for the spindle rotor in question
• 38. Regulation members for selective adaptation of the volume flow rate of the cooling fluid flow (9), managed by the control unit (25)
• 39. Vibration sensors to determine modified residual unbalance suggestions by different amounts of cooling fluid per spindle rotor internal cooling
• 40. Rotation unit per spindle rotor system, each fully assembled and balanced, primarily consisting of:
  • spindle rotor (2 and 3)
  • carrier shaft (4 and 5)
  • synchronisation toothing (8)
  • bearing, with (10) as fixed bearing plus working space shaft seals, for example with (13)
  • drive motor (18 and 19)
  • rotary encoder measurement system (20 and 21) thus, a total of two rotation units (40) per spindle compressor

Claims

1. A spindle compressor as a 2-shaft rotary positive-displacement machine, working without operating fluid in the working space, for conveying and compressing gaseous conveyed media, preferably steam, comprising a spindle rotor pair in a compressor housing (1) which has an inlet collection chamber (11) and an outlet collection chamber (12),

characterised in that the centre distance of the spindle rotor pair at the inlet-side end is at least 10% greater than at the outlet-side end, in that each of the two spindle rotors (2, 3) is driven by an electric motor (18, 19), and an electronic synchronisation controls the electric motors (18, 19), and in that the spindle rotors (2, 3) rotate contact-free.

2. The spindle compressor according to claim 1,

characterised in that one spindle rotor (2) has two teeth, in that the other spindle rotor (3) has three teeth, and in that the electronic synchronisation is a 2 to 3 synchronisation.

3. The spindle compressor according to claim 1,

characterised in that each spindle rotor (2 or 3) has an internal cooling means, which preferably is embodied as a cylindrical evaporator cooling bore (6) of radius RC2 on the 2-toothed spindle rotor (2) or of radius RC3 on the 3-toothed spindle rotor (3).

4. The spindle compressor according to claim 3,

characterised in that the evaporator cooling bore (6) has an inner structure with at least one of the following features, preferably more than one:

a) at least one cooling fluid guide groove (16), preferably with precise (deviation <1%) observance of the RC value, in particular with a.1) groove base faces with angles of inclination ψ(z) with 170°≤ψ(z)≤180° as f(z) and/or a.2) the outlet region has a larger surface for heat transfer than the inlet region,

b) cooling fluid distribution overflow grooves (17)

c) support points (7) for non-rotational support on the corresponding carrier shaft (4 or 5)

d) steam outlet (14) in the inlet chamber (11).

5. The spindle compressor according to claim 1,

characterised in that each spindle rotor system is embodied with the rotary unit (40) ready-assembled and balanced, and in that separator plates (26) are preferably provided for the final setting of the play between rotor heads and housing.

6. The spindle compressor according to claim 1,

characterised in that at least one vibration sensor (39) is provided and is connected to a control unit (25), and in that in the control unit (25) the supplied amount of the cooling fluid flow (9) is limited to the amount corresponding to a maximisation of the overall efficacy.

7. The spindle compressor according to claim 2,

characterised in that the critical bending speed of the 2-toothed spindle rotor is approximately (with a tolerance of preferably less than ±30%) 1.5 times higher than the critical bending speed of the 3-toothed spindle rotor (3).

8. The spindle compressor according to claim 1,

characterised in that the crossing angle alpha between the two spindle rotor axes of rotation in combination with the corresponding μ(z) value in the rotor longitudinal axis direction is such that, for each rotor, a cylindrical evaporator cooling bore (6) with minimal (that is to say appropriate for the particular tooth height in respect of material strength) wall thicknesses w is created on the supporting root-base body (32) (for example in accordance with the aforementioned position descriptions of E, S, V and L) under simultaneous consideration of the (preferably) blowhole-free profiling of the gas-conveying external thread (31) and critical bending speed “appropriate for the specific rotor spindle” and implementation of the inner volume ratio as iV value (as explained), wherein the gas-conveying external thread (31) is formed in the inlet region as a 2-toothed spindle rotor (2) preferably with cylindrical flattened portion (27).

9. The spindle compressor according to claim 1,

characterised in that the thermal situation for the working space components is regulated in an application-specific manner as basic step (as explained) during the component heat dissipation during operation to maintain the play values between avoidance of play reduction and excessive differences in the play values (as explained) and as FCT stage (as explained) during the component heat dissipation, to improve efficacy as diverted cooling fluid flow as separate cooling water flow via delayed evaporation

with cooling fluid injection (33) into the compressor working space, preferably in the region of the inlet collection chamber (11), which is all regulated and controlled by the control unit (25).

10. The spindle compressor according to claim 1,

characterised in that each spindle rotor (2, 3) consists of an aluminium alloy and is pressed on to a steel shaft (4, 5) at the support points (7) for conjoint rotation, and in that the gas-conveying external thread (31) is only then produced and the spindle rotor (2, 3) has an inner structure that is already completed.

11. The spindle compressor according to claim 1,

characterised in that the inner volume ratio is adapted to the current operating conditions via additional partial outlet openings (15).

12. The spindle compressor according to claim 1,

characterised in that a steam outlet (14) directly to the inlet is provided.

13. The spindle compressor according claim 1,

characterised in that a cylindrical flattened portion (27) is provided at the inlet of the 2-toothed spindle rotor, in particular in that the gas-conveying external thread (31) in the case of the 2-toothed spindle rotor (2) has the cylindrical flattened portion (27) in the inlet region.

14. The spindle compressor according claim 1,

characterised in that the 2-toothed spindle rotor (2) is provided with an intermediate support, whereby a weight reduction, in particular also for a lower moment of inertia during start-up (or braking) alongside high flexural rigidity, is preferably achieved for example from fibre-composite material suitable for vacuum, for example in the form of a CFRP material.

15. The spindle compressor according to claim 1,

characterised in that at least one cooling fluid feed (9.2 and 9.3) is provided, and in that each spindle rotor has a cylindrical evaporator cooling bore (6), which is connected to the cooling fluid feed (9.2 and 9.3).

16. The spindle compressor according to claim 1,

characterised in that each drive has a hollow shaft, in that the cooling fluid feed (9.2 and 9.3) to the cylindrical evaporator cooling bore (6) of a drive is provided through this hollow shaft, and in that the bearings (10) are preferably formed as durable bearings, in particular grease-lubricated-for life hybrid bearings, all-ceramic bearings, or also magnetic bearings.