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EP3134649B2 - Paire de rotors pour bloc de compresseur d'une visseuse - Google Patents
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EP3134649B2 - Paire de rotors pour bloc de compresseur d'une visseuse - Google Patents

Paire de rotors pour bloc de compresseur d'une visseuse Download PDF

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Publication number
EP3134649B2
EP3134649B2 EP15736405.0A EP15736405A EP3134649B2 EP 3134649 B2 EP3134649 B2 EP 3134649B2 EP 15736405 A EP15736405 A EP 15736405A EP 3134649 B2 EP3134649 B2 EP 3134649B2
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EP
European Patent Office
Prior art keywords
rotor
tooth
secondary rotor
profile
main
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Application number
EP15736405.0A
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German (de)
English (en)
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EP3134649A2 (fr
EP3134649B1 (fr
EP3134649B9 (fr
Inventor
Gerald WEIH
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Kaeser Kompressoren AG
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Kaeser Kompressoren AG
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Application filed by Kaeser Kompressoren AG filed Critical Kaeser Kompressoren AG
Priority to EP19190907.6A priority Critical patent/EP3597920B1/fr
Priority to EP18163593.9A priority patent/EP3358189B9/fr
Priority to EP23198449.3A priority patent/EP4273403A3/fr
Publication of EP3134649A2 publication Critical patent/EP3134649A2/fr
Publication of EP3134649B1 publication Critical patent/EP3134649B1/fr
Publication of EP3134649B9 publication Critical patent/EP3134649B9/fr
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
    • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C18/082—Details specially related to intermeshing engagement type pumps
    • F04C18/084—Toothed wheels
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01C—ROTARY-PISTON OR OSCILLATING-PISTON MACHINES OR ENGINES
    • F01C1/00—Rotary-piston machines or engines
    • F01C1/08—Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing
    • F01C1/082—Details specially related to intermeshing engagement type machines or engines
    • F01C1/084—Toothed wheels
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01C—ROTARY-PISTON OR OSCILLATING-PISTON MACHINES OR ENGINES
    • F01C1/00—Rotary-piston machines or engines
    • F01C1/08—Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing
    • F01C1/12—Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing of other than internal-axis type
    • F01C1/14—Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons
    • F01C1/16—Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons with helical teeth, e.g. chevron-shaped, screw type
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
    • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C18/12—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type
    • F04C18/14—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons
    • F04C18/16—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons with helical teeth, e.g. chevron-shaped, screw type
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
    • F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C18/12—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type
    • F04C18/14—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons
    • F04C18/20—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons with dissimilar tooth forms
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2240/00—Components
    • F04C2240/20—Rotors
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2240/00—Components
    • F04C2240/30—Casings or housings
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2240/00—Components
    • F04C2240/60—Shafts

Definitions

  • the invention relates to a pair of rotors for a compressor block of a screw machine, the pair of rotors consisting of a main rotor rotating about a first axis and a secondary rotor rotating about a second axis.
  • the invention also relates to a compressor block with a corresponding pair of rotors.
  • Screw machines whether as screw compressors or screw expanders, have been in practical use for several decades. Designed as screw compressors, they have replaced reciprocating compressors as compressors in many areas. With the principle of the interlocking pair of screws, not only gases can be compressed using a certain amount of work. The application as a vacuum pump also opens up the use of screw machines to achieve a vacuum. Finally, by passing pressurized gases through, a work output can also be generated the other way around, so that mechanical energy can also be obtained from pressurized gases using the principle of the screw machine.
  • Screw machines generally have two shafts arranged parallel to one another, on which a main rotor sits on the one hand and a secondary rotor on the other. Main rotor and secondary rotor engage with each other with appropriate helical gearing.
  • a compression space (working chambers) is formed by the tooth gap volumes between the teeth and a compressor housing in which the main and secondary rotors are accommodated. Starting from an intake area, the working chamber is initially closed and then continuously reduced in volume as the rotation of the main and secondary rotors progresses, so that the medium is compressed. Finally, as the rotation progresses, the working chamber opens towards a pressure window and the medium is pushed out into the pressure window. This process of internal compression distinguishes screw machines designed as screw compressors from Roots blowers, which work without internal compression.
  • a pair of rotors for a compressor block of a screw machine consisting of a secondary rotor rotating about a first axis and a main rotor rotating about a second axis, with the number of teeth on the main rotor being 3 and the number of teeth on the secondary rotor being 4, for example the US 2,622,787A known.
  • Typical pressure ratios can range from 1.1 to 20 depending on the tooth count ratio, where pressure ratio is the ratio of discharge pressure to intake pressure.
  • the compaction can take place in one or more stages.
  • Final pressures that can be achieved can be in the range from 1.1 bar to 20 bar, for example. To the extent that reference is made to pressure specifications in "bar" at this point or later in the present application, such pressure specifications relate to absolute pressures.
  • screw machines can also be used as compressors in different areas of technology.
  • a particularly preferred area of application is the compression of gases, such as air or inert gases (helium, nitrogen,).
  • gases such as air or inert gases (helium, nitrogen,).
  • a screw machine for compressing refrigerants, for example for air conditioning systems or refrigeration applications, although this places different structural requirements in particular.
  • fluid-injected compression in particular oil-injected compression, is usually used; but it is also possible to operate a screw machine according to the principle of dry compression.
  • screw compressors are sometimes also referred to as screw blowers.
  • the front section of the rotors in particular the front section of the secondary rotor, has a significant impact on energy efficiency.
  • the face section of the auxiliary rotor must find its equivalent in the face section of the main rotor.
  • the profile of the rotor in a plane perpendicular to the axis of the rotor is referred to as a face section.
  • Different types of face cut generation such as rotor-based or rack-based face cut generation methods, are now known from the prior art. Once you have decided on a specific method, a first draft face section is created in a first step. This is conventionally further optimized in several subsequent (revision) steps according to various criteria.
  • a short profile gap length should be combined with a small (pressure side) blowhole.
  • the two variables behave in opposite directions. This means that the smaller the blowhole is modeled, the larger the length of the profile gap will inevitably be. Conversely, the shorter the profile gap length, the larger the blowhole becomes. This is explained, for example, by Helpertz in his Dissertation "Method for the stochastic optimization of screw rotor profiles", Dortmund, 2003 on page 162 .
  • blow holes on the pressure side must not be made too large in order to minimize the backflow of already compressed medium into previous working chambers (i.e. in working chambers with lower pressure). Such backflows increase the energy expenditure for the total flow rate achieved and lead to an undesirable increase in the temperature and pressure level during compression, which reduces the overall efficiency.
  • the area of the blow hole (blow hole area) can be kept small by making the head curves of the profiles small in the front section. Specifically, this can be caused by a strong curvature in the tip area of the leading tooth flank of the secondary rotor and in the tip area of the trailing tooth flank of the main rotor. However, the greater this curvature, the more likely it is that you will end up in production-related border areas, since this leads, for example, to high wear on profile milling cutters and profile grinding wheels in the production of the main rotor and secondary rotor.
  • chamber gusset volume Another reason for efficiency-reducing internal leaks is the so-called chamber gusset volume, which can arise when the last working chamber (i.e. the working chamber in which the highest pressure prevails) is pushed out into the pressure window. From a certain angle of rotation of the rotors, the working chamber is no longer connected to the pressure window. A so-called chamber gusset volume remains between the two rotors and the pressure-side end wall of the housing.
  • This chamber gusset volume is disadvantageous because the enclosed compressed medium can no longer be pushed out into the pressure window, and as the rotors turn further it is compressed even further, which leads to unnecessarily high power consumption (for overcompression), an unnecessarily high additional heat input, noise development and a reduction the service life, in particular of the roller bearings of the rotors.
  • the specific performance is worsened by the fact that the portion enclosed in the chamber gusset volume returns to the suction side after overcompression and is therefore not available to the compressed air user.
  • oil-injected compressors there is also incompressible oil in the chamber gusset and is being squeezed.
  • Compact compressor blocks with a high utilization of construction volume are achieved through a large tooth gap volume, which in turn depends on the profile depth and the tooth thickness.
  • the object of the present invention consists in specifying a pair of rotors for a compressor block of a screw machine which has a high level of operational reliability and reasonable production costs characterized by smooth running and particular energy efficiency.
  • the rotor geometry is essentially characterized by the shape of the face section as well as by the rotor length and the angle of wrap, cf. " Method for the stochastic optimization of screw rotor profiles", thesis by Markus Helpertz, Dortmund, 2003, p. 11/12 .
  • the secondary rotor or main rotor has a predetermined, often different number of teeth of the same design per rotor.
  • the outermost circle drawn through the axis C1 or C2 over the crests of the teeth is referred to as the tip circle.
  • a base circle is defined in the face section by the points of the outer surface of the rotors closest to the axis.
  • the ribs are called the teeth of the rotor.
  • the grooves (or recesses) are accordingly referred to as tooth gaps.
  • the area of the tooth at and above the root circle defines the tooth profile.
  • the contour of the ribs defines the course of the tooth profile.
  • Base points F1 and F2 and a vertex F5 are defined for the tooth profile.
  • the vertex F5 or H5 is defined by the radially outermost point of the tooth profile. If the tooth profile has several points with the same maximum radial distance from the center point defined by the axis C1 or C2, i.e. the tooth profile follows an arc of a circle on the addendum circle at its radially outer end, then the vertex F5 lies exactly in the middle of this arc of a circle. A tooth gap is defined between two adjacent vertices F5.
  • the points radially closest to the axis C1 or C2 between a tooth under consideration and the respective adjacent tooth define base points F1 and F2.
  • the tooth profile at its lowest point follows the root circle in sections, the corresponding root point F1 or F2 then lies on half of this circular arc lying on the root circle .
  • a pitch circle is defined for both the secondary rotor and the main rotor.
  • screw machines as well as with gear wheels or friction wheels, there are always two circles in the front section of the toothing, which roll off each other during movement. These circles, on which in the present case the main rotor and the secondary rotor roll against each other, are referred to as respective pitch circles.
  • the pitch circle diameters of the main rotor and secondary rotor can be determined with the help of the center distance and the number of teeth ratio.
  • the circumferential speeds of the main rotor and secondary rotor are identical on the pitch circles.
  • tooth gap areas between the teeth and the respective addendum circle KK are defined, namely tooth gap area A6 between the profile of the secondary rotor NR between two adjacent vertices F5 and the addendum circle KK 1 or an area A7 as a tooth gap area between the profile of the main rotor (HR) between two neighboring vertices H5 and the tip circle KK 2 .
  • the tooth profile of the secondary rotor (but also of the main rotor) has a tooth flank that leads in the direction of rotation and a tooth flank that trails in the direction of rotation.
  • the leading tooth flank is referred to below as F V
  • the trailing tooth flank is F N .
  • the trailing tooth flank F N In its section between tip circle and root circle, the trailing tooth flank F N forms a point at which the curvature of the course of the tooth profile changes. This point is referred to below as F8 and divides the trailing tooth flank F N into a convexly curved portion between F8 and the addendum circle and a concavely curved portion between the root circle and F8. Small-scale profile changes, such as sealing strips or other local profile changes, are not taken into account when considering the change in curvature described above.
  • a wrap angle ⁇ is defined. This angle of wrap is the angle by which the face section is twisted from the suction-side to the pressure-side rotor face, cf. also the detailed explanations in connection with figure 8 .
  • the main rotor has a rotor length L HR , which is defined as the distance from a suction-side main rotor rotor face to a pressure-side main rotor rotor face.
  • the distance between the first axis C1 of the auxiliary rotor, which runs parallel to one another, and the second axis C2 of the main rotor is referred to below as the axis distance a.
  • the length of the main rotor L HR corresponds to the length of the secondary rotor L NR , with the length of the secondary rotor also being understood as the distance between a suction-side secondary rotor rotor face and a pressure-side secondary rotor rotor face.
  • a rotor length ratio L HR /a is defined, i.e. a ratio of the rotor length of the main rotor to the center distance.
  • the ratio L HR /a is a measure of the axial dimensioning of the rotor profile.
  • the line of action or the profile gap is created by the interaction of the main rotor and the secondary rotor with one another.
  • the line of action is as follows: the tooth flanks of the main rotor and secondary rotor touch each other at certain points when the teeth are free of play, depending on the rotational angle of the rotors. These points are called engagement points.
  • the geometric location of all points of action is called line of action and lets can already be calculated in two dimensions using the face section of the rotors, cf. Figure 7j .
  • the line of action is divided into two sections in the face section view by the connecting line between the two centers C1 and C2, specifically into a (comparatively short) suction-side section and a (comparatively long) pressure-side section.
  • the line of action can also be extended three-dimensionally and corresponds to the contact line of the main rotor and auxiliary rotor.
  • the axial projection of the three-dimensional line of action onto the front section plane results in turn from Figure 7j illustrated two-dimensional line of action.
  • the term "line of action” is used in the literature for both two-dimensional and three-dimensional considerations. In the following, unless otherwise stated, "line of action” is to be understood as meaning the two-dimensional line of action, ie the projection onto the face section.
  • the profile engagement gap is defined as follows: In the real compressor block of a screw machine, there is play between the two rotors when the main rotor and secondary rotor are installed axially apart.
  • the gap between the main rotor and the slave rotor is called the profile engagement gap and is the locus of all points where the two paired rotors touch each other or are closest to each other.
  • the profile engagement gap Through the profile engagement gap, the compressing and ejecting working chambers are connected to chambers that are still in contact with the suction side. The entire maximum pressure ratio is therefore present at the profile engagement gap. Already compressed working fluid is transported back to the suction side through the profile engagement gap and thus reduces the efficiency of the compression. Since the profile engagement gap would be the line of action in the case of backlash-free gearing, the profile engagement gap is also referred to as the “quasi line of action”.
  • Blowholes between working chambers are created by rounding the tips of the teeth of the profile.
  • the working chambers are connected to preceding and following working chambers via blowholes, so that (in contrast to the profile engagement gap) only the pressure difference from one working chamber to the next is present at a blowhole.
  • certain pairs of teeth are known to be common in screw machines, for example a rotor pair in which the main rotor has 3 and the auxiliary rotor has 4 teeth or a rotor pair in which the main rotor has 4 teeth and the auxiliary rotor has 5 teeth or furthermore a rotor pair geometry in which the main rotor has 5 teeth and the secondary rotor has 6 teeth.
  • Rotor pairs or screw machines with different tooth ratios may be used for different areas of application or purposes.
  • rotor pair arrangements with a tooth ratio of 4 / 5 main rotor with 4 teeth, secondary rotor with 5 teeth
  • the number of teeth or the number of teeth ratio specifies different types of rotor pairings and, as a result, also different types of screw machines, in particular screw compressors.
  • the ratio of the center distance ⁇ of the first axis C1 to the second axis C2 and the addendum circle radius rk is 1 a rk 1 set so that a rk 1 is at least 1.636 and at most 1.8, preferably at most 1.733, with the main rotor preferably being designed with a wrap angle ⁇ HR for which 240° ⁇ HR ⁇ 360° applies, and with the following preferably applying for a rotor length ratio L HR /a: 1.4 ⁇ L MR / a ⁇ 3.4 , where the rotor length ratio is formed from the ratio of the rotor length L HR of the main rotor and the center distance a and the rotor length L HR of the main rotor is formed by the distance between a suction-side main rotor rotor face and an opposite pressure-side main rotor rotor face.
  • the specified values for the rotor length ratio L HR /a and the wrap angle ⁇ HR represent advantageous or expedient values for the specified number of teeth ratio in order to define an advantageous rotor pairing in the axial dimension.
  • the aim is to combine a small blow hole with a short length of the profile engagement gap.
  • the two parameters behave in opposite directions, ie the smaller the blow hole is modeled, the larger the length of the profile contact gap will inevitably be. Conversely, the shorter the length of the profile-engaging gap, the larger the blowhole becomes.
  • a particularly favorable combination of the two parameters is achieved in the stressed areas. At the same time, a sufficiently high flexural rigidity of the secondary rotor is ensured.
  • a further preferred embodiment provides that in a face section consideration between the considered tooth of the secondary rotor (NR) and the respectively adjacent tooth of the secondary rotor, base points F1 and F2 on the root circle and on the radially outermost point of the tooth a vertex F5 are defined, with F1, F2 and F5, a triangle D Z is defined and in a radially outer region the tooth with its leading tooth flank F V formed between F5 and F2 has an area A1 and with its trailing tooth flank F N formed between F1 and F5 has an area A2 the triangle D Z protrudes and where 8 ⁇ A2/A1 ⁇ 60 is observed.
  • the partial tooth surface A1 on the leading tooth flank F V of the auxiliary rotor has a significant influence on the blow hole surface.
  • the partial tooth surface A2 on the trailing tooth flank F N of the auxiliary rotor has a significant influence on the length of the profile engagement gap, the chamber extension and the auxiliary rotor torque.
  • the pair of rotors has a secondary rotor, in which, in a cross-sectional view, base points F1 and F2 are defined between the tooth of the secondary rotor under consideration (NR) and the respective adjacent tooth of the secondary rotor, and an apex F5 is defined at the radially outermost point of the tooth.
  • a triangle D Z is defined by F1, F2 and F5 and where the leading tooth flank F V formed between F5 and F2 protrudes in a radially outer area of the tooth with an area A1 over the triangle D Z and in a radially inner area opposite the Triangle D Z with an area A3 recedes and where 7.0 ⁇ A3/A1 ⁇ 35 is observed.
  • base points F1 and F2 are defined between the tooth of the secondary rotor (NR) under consideration and the respective adjacent tooth of the secondary rotor (NR) and a vertex F5 is defined at the radially outermost point of the tooth are, wherein a triangle D Z is defined by F1, F2 and F5 and wherein the leading tooth flank F V formed between F5 and F2 protrudes in a radially outer region of the tooth with an area A1 over the triangle D Z , the tooth itself having a has a cross-sectional area A0 delimited by the arc of a circle B running between F1 and F2 around the center point defined by the axis C1, and where 0.5% ⁇ A1/A0 ⁇ 4.5% is observed.
  • Figures 7d as well as 7e.
  • a further preferred embodiment provides that in a face section consideration between the considered tooth of the secondary rotor (NR) and the respective adjacent tooth of the secondary rotor (NR) base points F1 and F2 and at the radially outermost point of the tooth an apex F5 are defined, with the between F1 and F2 arcs B running around the center point defined by the axis C1 defines a tooth pitch angle ⁇ corresponding to 360°/number of teeth of the secondary rotor (NR), with a point F11 being defined on half the arc B between F1 and F2, a point F11 being defined by a
  • the offset angle is preferably always positive, ie the offset is always in the direction of the direction of rotation and not in the opposite direction.
  • the tooth of the secondary rotor is curved towards the direction of rotation of the secondary rotor.
  • the offset should remain in the range specified as advantageous in order to enable a favorable compromise between the blowhole area, the shape of the line of action, the length and shape of the profile engagement gap, the auxiliary rotor torque, the bending stiffness of the rotors and the chamber extension into the pressure window.
  • Figure 7f referred.
  • trailing tooth flank F N of a tooth of the secondary rotor (NR) formed between F1 and F5 has a convex length portion of at least 45% to at most 95% in a face section consideration.
  • the secondary rotor is preferably designed in such a way that, in a front section view, the radial ray R drawn from the axis C1 of the secondary rotor (NR) through F5 divides the tooth profile into a surface portion A5 assigned to the leading tooth flank F V and a surface portion A4 assigned to the trailing tooth flank F N and whereby 5 ⁇ A 4 / A 5 ⁇ 14 is complied with. It should be pointed out once again at this point that the tooth profile is delimited radially inwards towards the axis C1 by the root circle FK1 .
  • the radial ray R divides the tooth profile in such a way that two disjoint surface portions with a total surface portion A5, which are assigned to the leading tooth flank F V , arise, cf. Figure 7g . If the vertex F5 were offset towards the leading tooth flank in such a way that the radial ray R not only touches the leading tooth flank F V but intersects it at two points, then two disjoint surface portions assigned to the leading tooth flank are defined with a total surface portion A5.
  • the area portion A4 assigned to the trailing tooth flank F N is then limited on the one hand by the radial ray R and in sections, namely between the two points of intersection of the leading tooth flank F V with the radial ray R, on the other hand also by the leading tooth flank F V .
  • a further preferred embodiment has a pair of rotors which is characterized in that the main rotor HR is designed with a wrap angle ⁇ HR for which the following applies: 290° ⁇ HR ⁇ 360°, preferably 320° ⁇ HR ⁇ 360°.
  • the pressure window area can be made larger with the same built-in volume ratio.
  • the axial extension of the working chamber to be pushed out, the so-called profile pocket depth is thereby also shortened. This reduces the exhaust throttling losses, especially at higher speeds, and thus enables better specific power.
  • an angle of wrap that is too large has a negative effect on the construction volume and leads to larger rotors.
  • a Bl denotes an absolute blowhole area on the pressure side
  • A6 and A7 denote tooth gap areas of the secondary rotor (NR) and the main rotor (HR), respectively, with the area A6 in a cross-sectional view being that between the profile of the secondary rotor (NR) between two adjacent vertices F5 and the addendum circle KK 1 enclosed area and the area A7 in a front section view the area enclosed between the profile of the main rotor (HR) between two adjacent vertices H5 and the addendum circle KK 2 .
  • ⁇ I designates a profile gap length factor, whereby the length of the profile engagement gap of a tooth gap is set in relation to the profile depth PT 1 . This allows a measure of the length of the profile engagement gap to be defined, regardless of the size of the screw machine.
  • the smaller the numerical value of the characteristic ⁇ I the shorter the profile gap of a tooth pitch with the same profile depth and thus the lower the leakage volume flow back to the suction side.
  • the goal of combining a small blow hole on the pressure side with a short profile gap results from the factor ⁇ I * ⁇ Bl .
  • the two key figures behave in opposite directions.
  • main rotor (HR) and secondary rotor (NR) are designed and coordinated in such a way that dry compression with a pressure ratio ⁇ of up to 3, in particular with a pressure ratio ⁇ of greater than 1 and up to 3 , is achievable, the pressure ratio being the ratio of the compression end pressure to the intake pressure.
  • a further preferred embodiment provides a pair of rotors such that the main rotor (HR) is designed to be operable with a peripheral speed in a range of 20 to 100 m/s in relation to a tip circle KK 2 .
  • a further embodiment has a pair of rotors which is characterized in that for a diameter ratio defined by the ratio of the tip circle radii of the main rotor (HR) and the secondary rotor (NR).
  • D v dk 2
  • a further embodiment has a pair of rotors which is characterized in that for a diameter ratio defined by the ratio of the tip circle radii of the main rotor (HR) and the secondary rotor (NR).
  • D v dk 2
  • the teeth of the secondary rotor taper outwards in a cross-sectional view, i.e. all circular arcs running perpendicularly to a radial ray emanating from the center point defined by the axis C1 and drawn through the point F5 from the trailing tooth flank F N to the leading tooth flank F V starting from F1 to F2 in the sequence radially outward decrease (or at least remain the same in sections).
  • the teeth of the secondary rotor are thus designed in such a way that there are no constrictions, ie the width of a tooth of the secondary rotor does not increase at any point, but instead decreases radially outwards or at most remains the same. This is considered useful in order to achieve a small blow hole on the pressure side with a nevertheless short profile engagement gap length.
  • the face section design of the secondary rotor is such that the direction of action of the torque, which results from a reference pressure on the partial surface of the secondary rotor delimiting a working chamber, is directed counter to the direction of rotation of the secondary rotor.
  • Such an end section design has the effect that the entire torque from the gas forces on the secondary rotor is directed counter to the direction of rotation of the secondary rotor. This achieves a defined flank contact between the trailing secondary rotor flank F N and the leading main rotor flank. This contributes to avoiding the problem of so-called rotor rattling, which can occur in unfavorable, in particular non-stationary, operating situations.
  • Rotor rattling is understood to mean a leading and lagging of the secondary rotor about its axis of rotation superimposed on the uniform rotational movement, which is accompanied by a rapidly alternating impact of the trailing secondary rotor flanks on the leading main rotor flanks and then the leading secondary rotor flanks on the trailing main rotor flanks, etc.
  • This problem occurs in particular when the moment from the gas forces together with other moments (eg from bearing friction) on the secondary rotor is undefined (eg close to zero), which is effectively avoided by the advantageous face section design.
  • the main rotor (HR) and the secondary rotor (NR) are designed to convey air or inert gases, such as helium or nitrogen, and are matched to one another.
  • the profile of a tooth of the secondary rotor is asymmetrically formed in a face section view with respect to the radial ray R drawn from the center point, which is defined by the axis C1, through the apex F5.
  • the leading tooth flank and trailing tooth flank of each tooth are therefore asymmetrical in relation to one another.
  • This asymmetrical design is already known per se for screw compressors. However, it makes a significant contribution to efficient compaction.
  • a further preferred embodiment provides that a point C on the connecting section is viewed in a front section C1 C2 is defined between the first axis C1 and the second axis C2, where the pitch circles WK 1 of the secondary rotor (NR) and WK 2 of the main rotor (HR) touch, that K5 is the intersection of the root circle FK 1 of the secondary rotor (NR) with the connecting section C1 C2 defined, where r 1 measures the distance between K5 and C, and that K4 denotes the point of the suction-side part of the line of action furthest from the connecting line C1 C2 is spaced between C1 and C2, where r 2 measures the distance between K4 and C and where: 0.9 ⁇ right 1 right 2 ⁇ 0.875 ⁇ e.g 1 e.g 2 + 0.22 with z 1 : number of teeth on the secondary rotor (NR) and z 2 : number of teeth on the main rotor (HR).
  • the pair of rotors is designed and designed in such a way that for a rotor length ratio L HR /a applies: 0.85 * e.g 1 / e.g 2 + 0.67 ⁇ L MR / a ⁇ 1.26 * e.g 1 / e.g 2 + 1:18 , preferred 0.89 * e.g 1 / e.g 2 + 0.94 ⁇ L MR / a ⁇ 1.05 * e.g 1 / e.g 2 + 1.22 , with z 1 : number of teeth on the secondary rotor (NR) and z 2 : number of teeth on the main rotor (HR), with the rotor length ratio L HR /a indicating the ratio of the rotor length L HR to the center distance a and rotor length L HR the distance between the suction side Main rotor rotor face to the pressure-side main rotor rotor face.
  • the flexural rigidity of the rotors is sufficiently high so that the rotors do not deflect significantly during operation and the gaps (between the rotors or between the rotors and the compressor housing) can therefore be made relatively narrow without the risk of that under unfavorable operating conditions (high temperatures and/or high pressures) the rotors collide or collide in the compressor housing.
  • Narrow gaps offer the advantage of low backflow and thus contribute to energy efficiency. At the same time, operational safety is guaranteed despite the small gap dimensions.
  • a high flexural rigidity of the rotors is also advantageous in rotor production in order to comply with the high requirements for shape tolerances.
  • the ratio of L HR /a is dimensioned so large that the center distance a is not excessively large in relation to the rotor length L HR .
  • the gap lengths can be kept small; this reduces the backflow into the previous working chambers and this in turn improves energy efficiency.
  • the axial forces resulting from the pressurized, pressure-side end faces of the rotors can also be kept small by small-dimensioned end faces. These axial forces act on the rotors and in particular on the rotor bearings during operation. By minimizing these axial forces, the load on the (roller) bearings can be minimized or the bearings can be dimensioned smaller.
  • the configuration of the tooth profile of the secondary rotor described above is relevant above all for a tooth number ratio of 3/4 or 4/5. With such a tooth number ratio, the blowhole area can be reduced by complying with the condition given above. With a tooth number ratio of 5/6, however, an aforementioned point of contact or aforementioned points of intersection with the leading tooth flank F V does not appear to be desirable, since the teeth of the secondary rotor may then be too thin and consequently too flexible.
  • a compressor block comprising a compressor housing and a pair of rotors as described above is claimed as being according to the invention, the pair of rotors comprising a main rotor HR and a secondary rotor NR, which are each rotatably mounted in the compressor housing.
  • the compressor block is designed in such a way that the front section is designed in such a way that the working chamber formed between the tooth profiles of the main rotor (HR) and secondary rotor (NR) can be pushed out essentially completely into the pressure window.
  • the front section design of the two rotors advantageously ensures that no chamber gusset volume forms between the two rotors when the working chamber is pushed out into the pressure window.
  • the compression can be carried out particularly efficiently, since there is no return flow of medium that has already been compressed to the intake side, and this also means that there is no additional heat input.
  • the entire compressed volume can be used by downstream compressed air consumers.
  • a shaft end of the main rotor is guided out of the compressor housing and designed for connection to a drive, with both shaft ends of the secondary rotor preferably being completely accommodated within the compressor housing.
  • Example 1 Comparative example 1 Comparative example 2 Number of teeth HR z 2 3 3 4 5 Number of teeth NR z 1 4 4 5 6 PTrel [-] 0.588 0.54 0.528 0.455 a/rk 1 [-] 1.66 1.72 1,764 1.78
  • the profiles were created with the following center distances a: Example 1 Example 2 Comparative example 1 Comparative example 2 Center distance a[mm] 127 111 This results in the following cross section main dimensions: Example 1 Example 2 Comparative example 1 Comparative example 2 Dk 2 [mm] 191 186.1 186 154 Dk 1 [mm] 153 147.7 144 124.7 bw 2 [mm] 54.4 56.4 50.5 bw 1 [mm] 72.6 70.6 60.5 Other main dimensions of the rotors: Example 1 Example 2 Comparative example 1 Comparative example 2 Rotor length L HR [mm
  • Radial R 10 Radial ray R 10 has 2 points of intersection with the leading tooth flank F v Area ratio A4/A5 [-] 7.5 10.1 5.5 - wrap angle ⁇ HR 334 7° 330.3 330.3 ⁇ Bl [%] 0.159 0.086 0.106 0.18 ⁇ Bl* ⁇ l [%] 0.94 0.53 0.631 1,058
  • Profile front section design regarding chamber extension The working chamber can essentially be completely pushed out into the pressure window.
  • Profile face section design with regard to secondary rotor torque The effective direction of the NR torque resulting from the gas forces is directed against the direction of rotation of the secondary rotor.
  • the isentropic block efficiency compared to the prior art is for the second embodiment to 3/4 tooth number ratio in figure 5 illustrated. Two curves of the same pressure ratio are shown there. The pressure ratio actually shown is 2.0 (ratio of outlet pressure to inlet pressure). The isentropic block efficiency has been significantly improved compared to the values achievable with the prior art.
  • the in the figures 5 and 6 The delivery quantity specified in each case corresponds to the delivery volume flow of the compressor block in relation to the suction condition.
  • direction of rotation 24 of the secondary rotor and the necessarily resulting direction of rotation of the main rotor when operating as a compressor.
  • the leading tooth flank F V and the trailing tooth flank F N are marked on a secondary rotor tooth.
  • a tooth gap 23 is marked as representative of all tooth gaps of the secondary rotor.
  • the based on Figure 7a shown profile course of the leading tooth flank F V and the trailing tooth flank F N corresponds to the basis of figure 4 for a non-inventive tooth ratio of 5/6 explained comparative example.
  • Figure 7b shows the tooth gap surfaces A6 and A7 as well as a side view of a blowhole in a cross-sectional view.
  • the profile curves shown to explain the tooth gap areas A6 and A7 correspond to that for a tooth number ratio of 3/4 based on figure 1 illustrated embodiment.
  • the coordinate system is spanned by the u-axis parallel to the front faces of the rotor along the pressure-side intersection edge 11.
  • the blow hole on the pressure side lies in the coordinate system described and quite specifically in a plane perpendicular to the rotor end faces between the pressure side intersection edge 11 and a line of action point K2 of the part of the line of action on the pressure side.
  • the line of action 10 is divided into two sections by the connecting line between the two centers C1 and C2: the suction-side part of the line of action is shown below, the pressure-side part above the connecting line.
  • K2 designates the point of the pressure-side part of the line of action 10 which is spaced farthest from the straight line through C1 and C2.
  • the intersection of the tip circles of the two rotors creates an intersection edge 11 on the pressure side and an intersection edge 12 on the suction side Figure 7b the pressure-side intersection edge 11 is shown as a point in a cross-sectional view. The same applies to the representation of the suction-side intersection edge 12.
  • the u-axis is parallel to the rotor faces and corresponds to the vector from the line of action point K2 to the pressure-side intersection edge 11 in a face section view. Further details on the blow hole area A Bl on the pressure side can be found in Figure 7k .
  • Figure 7c shows a front section view of a tooth of the auxiliary rotor with a number of teeth not according to the invention with the concentric circular arcs B 25 , B 50 , B 75 running inside the rotor tooth around the center point C1 with the associated radii r 25 , r 50 , r 75 and the associated arc lengths b 25 , b50 , b75 .
  • the circular arcs B 25 , B 50 , B 75 are each delimited by the leading tooth flank F V and the trailing tooth flank F N .
  • the based on Figure 7c shown profile course of the leading tooth flank F V and the trailing tooth flank F N corresponds to the basis of figure 4 for a gear ratio of 5/6 explained embodiment.
  • Figure 7d shows in a cross-sectional view between the considered tooth of the secondary rotor and the respectively adjacent tooth of the secondary rotor base points F1 and F2 on the root circle and a vertex F5 at the radially outermost point of the tooth. Furthermore, the triangle D Z defined by the points F1, F2 and F5 is shown, with the secondary rotor being shown with a number of teeth that is not according to the invention.
  • tooth pitch angle ⁇ corresponding to 360°/number of teeth of the secondary rotor.
  • profile course of the leading tooth flank F V and the trailing tooth flank F N corresponds to the basis of figure 4 for a gear ratio of 5/6 explained embodiment.
  • Figure 7e shows the cross-sectional area A0 of a tooth of the secondary rotor, which is delimited by the circular arc B running between F1 and F2 around the center point C1, in a cross-sectional view.
  • the based on Figure 7e shown profile course of the leading tooth flank F V and the trailing tooth flank F N corresponds to the basis of figure 4 for a non-inventive tooth ratio of 5/6 explained comparative example.
  • Figure 7f shows the offset angle ⁇ in a front section view of a secondary rotor, which is shown with a number of teeth that is not according to the invention. This is defined by the offset from point F11 to point F12 viewed in the direction of rotation of the secondary rotor.
  • F11 is a point on half the circular arc B between F1 and F2 around the center C1 and therefore corresponds to the intersection of the bisector of the tooth pitch angle ⁇ with the circular arc B.
  • F12 results from the intersection of the radial ray R drawn from the center C1 to the vertex F5 with the arc of a circle B.
  • the basis of Figure 7f shown profile course of the leading tooth flank F V and the trailing tooth flank F N corresponds to the basis of figure 4 for a non-inventive tooth ratio of 5/6 explained comparative example.
  • Figure 7g shows the inflection point F8 on the trailing tooth flank F N of the secondary rotor, in which the curvature of the course of the tooth profile changes between the addendum and root circle, in a cross-sectional view.
  • the trailing tooth flank F N of the auxiliary rotor is divided by point F8 into a substantially convexly curved portion between F8 and the apex F5 and a substantially concavely curved portion between F8 and the base point F1.
  • Figure 7h shows two points of intersection of the radial ray R 10 from C1 to F10 with the leading tooth flank F V of the secondary rotor in a cross-sectional view, with point F10 designating that point on the leading tooth flank F V which lies on the addendum circle KK 1 with rk 1 and is furthest from F5 is spaced.
  • the tooth flank follows a circular arc ARC 1 with radius rk 1 radially on the outside over a defined section around the center point of the auxiliary rotor defined by the axis C1.
  • the profile curves of the leading tooth flank F V and the trailing tooth flank F N explained above correspond to the exemplary embodiment described for a tooth number ratio of 3/4 figure 1 .
  • Figure 7i shows the tooth profile divided by the radial ray R drawn from C1 to F5 in a face section view.
  • the tooth profile is divided into a surface portion A4 associated with the trailing tooth flank F N and a surface portion A5 associated with the leading tooth flank F V .
  • the based on Figure 7i The profile curves of the leading tooth flank F V and the trailing tooth flank F N explained above correspond to the comparative example described for a non-inventive tooth number ratio of 5/6 figure 4 .
  • Figure 7j shows a cross-sectional view of the line of action 10 between the main and auxiliary rotor with a non-inventive tooth ratio and the two concentric circles around the point C with the radii r 1 and r 2 to describe the characteristic features of the course of the suction-side part of the line of action.
  • the line of action 10 is divided into two sections by the connecting distance between the first axis C1 and the second axis C2: the suction-side part of the line of action is below, the pressure-side part is above the connecting distance C1 C2 shown.
  • Point C is the point of contact of pitch circle WK 1 of the secondary rotor with pitch circle WK 2 of the main rotor.
  • K4 designates the point of the suction-side part of the line of action which is farthest from the connecting section between C1 and C2.
  • Radius r 1 is the distance between K5 and C, radius r 2 denotes the distance between K4 and C.
  • Figure 7k shows a pressure-side blow hole area A Bl of a working chamber, specifically in a sectional view perpendicular to the rotor end faces.
  • the delimitation of the blow hole area A Bl arises from the intersection line 27 of the imaginary flat surface described above with the leading secondary rotor flank F V , the intersection line 26 of the plane with the trailing HR flank and a straight section [K1 K3] of the pressure-side intersection edge 11.
  • figure 9 shows a schematic sectional view of a compressor block 19 comprising a housing 15 and mounted therein two rotors geared to one another in pairs, namely a main rotor HR and a secondary rotor NR.
  • the main rotor HR and the secondary rotor NR are each rotatably mounted in the housing 15 via suitable bearings 16 .
  • a drive power can be applied to a shaft 17 of the main rotor HR, for example with a motor (not shown) via a clutch 18 .
  • the compressor block shown is an oil-injected screw compressor, in which the torque is transmitted between the main rotor HR and the secondary rotor NR directly via the rotor flanks.
  • a synchronization gear (not shown).
  • an intake port for sucking in the medium to be compressed and an outlet for the compressed medium are also not shown.
  • FIG 10 a main rotor HR interlocked with one another and a secondary rotor NR are also shown in a perspective view.
  • figure 11 shows the spatial line of action 10 of exactly one tooth gap 23.
  • the profile gap length I sp is the length of the spatial line of action of exactly one tooth gap 23. This therefore corresponds to the profile gap length of exactly one tooth pitch.
  • the total torque from the gas forces on the secondary rotor is made up of the sum of the torque effects of the gas pressures in all working chambers on the sub-surfaces of the secondary rotor that delimit the respective working chambers.
  • a partial surface (22) of the secondary rotor delimiting a working chamber is shown hatched as an example.
  • Figure 12b shows the division of the in Figure 12a partial surface (22) delimiting a working chamber into an area (28) shown with dots and an area (29) shown with cross-hatching. Only the cross-hatched area (29) contributes to the torque.
  • the partial surface (22) results from the specific face section design and the pitch of the secondary rotor.
  • the pitch of the slave rotor refers to the pitch of the helical splines of the slave rotor.
  • the three-dimensional line of action (10) which is also shown and delimits the partial surface is also defined by the design of the face section of the secondary rotor and the pitch.
  • Partial surface (22) is also bounded by intersection line (27). Details of cutting line (27) have already been included in the figures 7b and 7k shown and described. The same applies to the line of action point K2.
  • the effective direction of the torque that the gas pressure in the working chamber (or any reference pressure) causes on the partial surface of the secondary rotor that delimits the working chamber is determined by the face section design of the secondary rotor.
  • the above-described advantageous front section design of the secondary rotor (NR) therefore leads to an effective direction (25) of the torque from the gas forces for each partial surface (22) of the secondary rotor that delimits a working chamber and thus for the entire secondary rotor, which is opposite to the direction of rotation (24) of the secondary rotor is directed, whereby the rotor clatter is effectively avoided.
  • the illustrated exemplary embodiments prove that with the present invention a considerable increase in efficiency could be achieved for a pair of rotors used in screw machines, consisting of a main rotor and a secondary rotor with a corresponding profile geometry.
  • profile profiles can also be generated using publicly accessible computer programs—as is well known to those skilled in the art.
  • SV_Win a project of the Vienna University of Technology
  • DISCO software in particular the SCORPATH module from City University London (Centre for Positive Displacement Compressor Technology). General information on this can be found at http://www.city-compressors.co. UK/. Information on installing the software can be found at http://www.staff.city.ac.uk/ ⁇ ra600/DISCO/DISCO/Installation%20instructions.pdf.
  • a preview of the DISCO software can be found at http://www.staff.city.ac.uk/ ⁇ ra600/DISCO/DISCO%20Preview.htm.
  • ScrewView software is also available in the Dissertation “Directed Evolutionary Algorithms” by Stefan Berlik, Dortmund 2006 (p. 173 f .) mentioned.
  • the ScrewView software is described in more detail on the website http://pi.informatik.unirisonen.de/Employees/berlik/projects/ in connection with the project "Method for designing dry-running rotary displacement machines”.
  • a tooth with a trailing rotor flank FN and a leading rotor flank F V is specifically generated as follows, with the Figures 15 and 16 Comparative examples with a tooth number ratio not according to the invention show:
  • the section S1 to S2 results from a circular arc on the secondary rotor NR around the center C1, generated by the circular arc section T1 to T2 around the center C2 on the main rotor HR.
  • the section S2 to S3 results from an envelope curve to a trochoid, generated by arc section T2 to T3 around the center point M4 on the main rotor HR.
  • the section S3 to S4 is defined by an arc of a circle around the center point M1.
  • the section S4 to S5 is defined by an arc of a circle around the center point M2.
  • the section S5 to S6 is defined by an arc of a circle around the center point C1.
  • the subsequent section S6 to S7 is specified by an arc of a circle around the center point M3.
  • the section S7 to S1 is specified by an envelope curve to form a trochoid, generated by the circular arc section T7 to T1 around the center point M5 on the main rotor HR.
  • the profile of the teeth of the main rotor HR is for the embodiment according to Figures 1 and 2 and the comparative examples according to FIG Figures 3 to 4 also based on the Figures 13 to 16 briefly explained below, where Figures 15 and 16 show comparative examples not according to the invention.
  • the section T1-T2 results from an arc of a circle on the main rotor HR around the center point C2 on the main rotor HR.
  • the section T2-T3 is defined by the circular arc on the main rotor HR around the center point M4.
  • Section T3-T4 results from an envelope curve to a trochoid generated by section S3-S4 on slave rotor NR.
  • Section T4-T5 is defined by an envelope curve to a trochoid generated by section S4-S5 on the auxiliary rotor.
  • the section T5-T6 is defined by a circular arc around the center C2 generated by the circular arc section S5-S6 around the center C1 on the slave rotor NR.
  • Section T6-T7 results from an envelope curve to form a trochoid, generated by section S6-S7 on secondary rotor NR.
  • the section T7-T1 is defined by an arc of a circle around the center point M5.
  • the profile curves of the secondary rotor NR and the main rotor HR are of course coordinated with one another and insofar as the envelope curves of a trochoid correspond to circular arc sections on the counter-rotor.
  • a tangential transition from one section to the next is guaranteed.
  • a general procedure for calculating the profile of the counter-rotor is, for example, in Dissertation by Helpertz, "Method for the stochastic optimization of screw rotor profiles", Dortmund, 2003, p. 60 ff . described.

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Claims (24)

  1. Paire de rotors pour un bloc compresseur d'une machine à vis, sachant que la paire de rotors est composée d'un rotor secondaire (NR) tournant autour d'un premier axe (C1) et d'un rotor principal (HR) tournant autour d'un deuxième axe (C2), sachant que le nombre de dents (z2) pour le rotor principal (HR) est de 3 et le nombre de dents (z1) pour le rotor secondaire (NR) est de 4,
    caractérisée en ce que
    la profondeur de profil relative du rotor secondaire PT rel = rk 1 − rf 1 rk 1
    Figure imgb0049
     est d'au moins 0,5, de préférence d'au moins 0,515, et d'au plus 0,595, sachant que rk1 est un rayon de cercle de tête tracé autour de la circonférence extérieure du rotor secondaire (NR) et rf1 est un rayon de cercle de base partant de la base de profil du rotor secondaire,
    sachant que le rapport de l'entraxe α du premier axe (C1) au deuxième axe (C2) et du rayon de cercle de tête rk1 a rk 1
    Figure imgb0050
     est d'au moins 1,636 et d'au plus 1,8, de préférence d'au plus 1,733,
    sachant que le rotor principal est de préférence constitué avec un angle d'enveloppement ΦHR conforme à 240° ≤ ΦHR ≤ 360°, et sachant qu'un rapport de longueur de rotor LHR/α est conforme à : 1,4 ≤ L HR / a ≤ 3,4 ,
    Figure imgb0051
    sachant que le rapport de longueur de rotor est formé à partir du rapport de la longueur de rotor LHR du rotor principal et de l'entraxe α, et la longueur de rotor LHR du rotor principal est formée par la distance d'une face frontale de rotor principal côté aspiration à une face frontale de rotor principal opposée côté refoulement.
  2. Paire de rotors selon la revendication 1, caractérisée en ce que, dans une vue en coupe frontale, des arcs de cercle B25, B50, B75 passant à l'intérieur d'une dent de rotor secondaire, dont le centre commun est donné par l'axe C1, sont définis, sachant que le rayon r25 de B25 a la valeur r25 = rf1+0,25∗(rk1-rf1), le rayon r50 de B50 a la valeur r50 = rf1+0,5∗(rk1-rf1), et le rayon r75 de B75 a la valeur rf1+0,75∗(rk1-rf1), et sachant que les arcs de cercle B25, B50, B75 sont délimités respectivement par le flanc de dent d'attaque FV et le flanc de dent de fuite FN,
    sachant que des rapports d'épaisseur de dent sont définis comme rapports des longueurs d'arc b25, b50, b75 des arcs de cercle B25, B50, B75 avec ε1 = b50/b25 et ε2 = b75/b25 et le dimensionnement suivant est respecté :
    0,65 ≤ ε1 < 1,0 et/ou 0,50 ≤ ε2 ≤ 0,85, de préférence 0,80 ≤ ε1 < 1,0 et/ou 0,50 ≤ ε2 ≤ 0,79.
  3. Paire de rotors selon la revendication 1 ou 2, caractérisée en ce que, dans une vue en coupe frontale, des bases F1 et F2 sont définies entre la dent considérée du rotor secondaire (NR) et la dent respectivement voisine du rotor secondaire et un sommet F5 est défini au point radialement le plus extérieur de la dent,
    sachant qu'un triangle DZ est défini par F1, F2 et F5, et
    sachant que dans une zone radialement extérieure, la dent dépasse du triangle DZ à raison d'une aire A1 par son flanc de dent d'attaque Fv constitué entre F5 et F2 et à raison d'une aire A2 par son flanc de dent de fuite FN constitué entre F1 et F5 et sachant que 8 ≤ A2/A1 ≤ 60 est respecté.
  4. Paire de rotors selon l'une des revendications 1 à 3, caractérisée en ce que, dans une vue en coupe frontale, des bases F1 et F2 sont définies entre la dent considérée du rotor secondaire (NR) et la dent respectivement voisine du rotor secondaire et un sommet F5 est défini au point radialement le plus extérieur de la dent,
    sachant qu'un triangle DZ est défini par F1, F2 et F5, et
    sachant que le flanc de dent d'attaque FV constitué entre F5 et F2 dépasse du triangle Dz à raison d'une aire A1 dans une zone radialement extérieure de la dent et est en retrait par rapport au triangle Dz à raison d'une aire A3 dans une zone radialement intérieure et sachant que 7,0 ≤ A3/A1 ≤ 35 est respecté.
  5. Paire de rotors selon l'une des revendications 1 à 4, caractérisée en ce que, dans une vue en coupe frontale, des bases F1 et F2 sont définies entre la dent considérée du rotor secondaire (NR) et la dent respectivement voisine du rotor secondaire (NR) et un sommet F5 est défini au point radialement le plus extérieur de la dent,
    sachant qu'un triangle DZ est défini par F1, F2 et F5, et
    sachant que le flanc de dent d'attaque FV constitué entre F5 et F2 dépasse du triangle Dz à raison d'une aire A1 dans une zone radialement extérieure de la dent,
    sachant que la dent elle-même présente une aire de coupe transversale A0 délimitée par l'arc de cercle B passant entre F1 et F2 autour du centre défini par l'axe C1 et sachant que 0,5 % ≤ A1/A0 ≤ 4,5 % est respecté.
  6. Paire de rotors selon l'une des revendications 1 à 5, caractérisée en ce que, dans une vue en coupe frontale, des bases F1 et F2 sont définies entre la dent considérée du rotor secondaire (NR) et la dent respectivement voisine du rotor secondaire (NR) et un sommet F5 est défini au point radialement le plus extérieur de la dent,
    sachant que l'arc de cercle B passant entre F1 et F2 autour du centre défini par l'axe C1 définit un angle de pas de dent γ correspondant à 360°/nombre de dents du rotor secondaire (NR),
    sachant qu'un point F11 est défini sur le demi-arc de cercle B entre F1 et F2, sachant qu'une projection radiale R tracée depuis le centre défini par l'axe C1 du rotor secondaire (NR) par le sommet F5 croise l'arc de cercle B en un point F12, sachant qu'un angle de décalage β est défini par le décalage de F11 par rapport à F12 considéré dans le sens de rotation du rotor secondaire (NR), et sachant que 14 % ≤ δ ≤ 25 %
    Figure imgb0052
     est respecté, avec δ = β γ ∗ 100 %
    Figure imgb0053
    .
  7. Paire de rotors selon l'une des revendications 1 à 6, caractérisée en ce que, dans une vue en coupe frontale, le flanc de dent de fuite FN constitué entre F1 et F5 d'une dent du rotor secondaire (NR) présente une part de longueur convexe d'au moins 45 % à au plus 95 %.
  8. Paire de rotors selon l'une des revendications 1 à 7, caractérisée en ce que, dans une vue en coupe frontale, la projection radiale tracée depuis l'axe C1 du rotor secondaire (NR) par F5 divise le profil de dent en une part d'aire A5 associée au flanc de dent d'attaque Fv et une part d'aire A4 associée au flanc de dent de fuite FN, et sachant que 5 ≤ A 4 / A 5 ≤ 14
    Figure imgb0054
     est respecté.
  9. Paire de rotors selon l'une des revendications 1 à 8, caractérisée en ce que le rotor principal (HR) est constitué avec un angle d'enveloppement ΦHR conforme à 290° ≤ ΦHP ≤ 360°, de préférence 320° ≤ ΦHR ≤ 360°.
  10. Paire de rotors selon l'une des revendications 1 à 9, caractérisée en ce qu'un facteur de trou de soufflage µBl est d'au moins 0,02 % et d'au plus 0,4 %, de préférence d'au plus 0,25 %,
    sachant que
    Figure imgb0055
     et
    sachant que ABl désigne une aire absolue de trou de soufflage côté refoulement et A6 et A7 désignent respectivement des aires d'entre-dents du rotor secondaire (NR) et du rotor principal (HR), sachant que l'aire A6, dans une vue en coupe frontale, désigne l'aire comprise entre le tracé de profil du rotor secondaire (NR) entre deux sommets F5 voisins et le cercle de tête KK1, et l'aire A7, dans une vue en coupe frontale, désigne l'aire comprise entre le tracé de profil du rotor principal (HR) entre deux sommets H5 voisins et le cercle de tête KK2.
  11. Paire de rotors selon l'une des revendications 1 à 10, caractérisée en ce que, pour un facteur trou de soufflage / longueur d'entrefer µl ∗ µBl, 0,1 % ≤ μ I * μ BI ≤ 1,72 %
    Figure imgb0056
     est respecté, avec
    Figure imgb0057
    sachant que Isp désigne la longueur de l'entrefer d'un entre-dent du rotor secondaire et PT1 désigne la profondeur de profil du rotor secondaire avec PT1 = rk1 - rf1,
    et
    Figure imgb0058
    sachant que ABI désigne une aire absolue de trou de soufflage et A6 et A7 désignent respectivement des aires de profil du rotor secondaire (NR) et du rotor principal (HR), sachant que l'aire A6, dans une vue en coupe frontale, désigne l'aire comprise entre le tracé de profil du rotor secondaire (NR) entre deux sommets F5 voisins et le cercle de tête KK1, et l'aire A7, dans une vue en coupe frontale, désigne l'aire comprise entre le tracé de profil du rotor principal (HR) entre deux sommets H5 voisins et le cercle de tête KK2.
  12. Paire de rotors selon l'une des revendications 1 à 11, caractérisée en ce que le rotor principal (HR) et le rotor secondaire (NR) sont constitués et adaptés l'un à l'autre de telle façon qu'une compression sèche soit réalisable avec un rapport de pression Π allant jusqu'à 3, en particulier avec un rapport de pression Π supérieur à 1 et allant jusqu'à 3, sachant que le rapport de pression est le rapport de la pression finale de compression à la pression d'aspiration.
  13. Paire de rotors selon l'une des revendications 1 à 12, caractérisée en ce que le rotor principal (HR), relativement à un cercle de tête KK2, est constitué pour pouvoir fonctionner à une vitesse circonférentielle comprise dans une plage de 20 à 100 m/s.
  14. Paire de rotors selon l'une des revendications 1 to 13, caractérisée en ce que, pour un rapport de diamètre défini par le rapport des rayons de cercle de tête du rotor principal (HR) et du rotor secondaire (NR), D v = Dk 2 Dk 1 = rk 2 rk 1
    Figure imgb0059
     1,145 ≤ D V ≤ 1,30
    Figure imgb0060
     est respecté, sachant que Dk1 désigne le diamètre du cercle de tête KK1 du rotor secondaire (NR) et Dk2 désigne le diamètre du cercle de tête KK2 du rotor principal (HR).
  15. Paire de rotors selon l'une des revendications 1 à 14, caractérisée en ce que, dans une vue en coupe frontale, les longueurs d'arc b(r) passant à l'intérieur d'une dent du rotor secondaire des arcs de cercle concentriques respectivement associés présentant le rayon rf1 < r < rk1 et le centre commun défini par l'axe C1 sont délimités respectivement par le flanc de dent d'attaque Fv et le flanc de dent de fuite FN et les longueurs d'arc b(r) diminuent de façon monotone au fur et à mesure que le rayon r augmente.
  16. Paire de rotors selon l'une des revendications 1 à 15, caractérisée en ce que la configuration en coupe frontale du rotor secondaire (NR) est réalisée de telle façon que le sens d'action du couple qui résulte d'une pression de référence sur la surface partielle du rotor secondaire qui délimite une chambre de travail est contraire au sens de rotation du rotor secondaire.
  17. Paire de rotors selon l'une des revendications 1 à 16, caractérisée en ce que le rotor principal (HR) et le rotor secondaire (NR) sont constitués et adaptés l'un à l'autre pour le refoulement d'air ou de gaz inertes tels que l'hélium ou l'azote.
  18. Paire de rotors selon l'une des revendications 1 à 17, caractérisée en ce que, dans une vue en coupe frontale, le profil d'une dent du rotor secondaire est constitué de façon asymétrique relativement à la projection radiale R tracée depuis le centre qui est défini par l'axe C1 par le sommet F5.
  19. Paire de rotors selon l'une des revendications 1 à 18, caractérisée en ce que, dans une vue en coupe frontale, un point C est défini sur le trajet de liaison ( C1C2) entre le premier axe (C1) et le deuxième axe (C2), où les cercles primitifs WK1 du rotor secondaire (NR) et WK2 du rotor principal (HR) se touchent, en ce que K5 définit le point d'intersection du cercle de base FK1 du rotor secondaire (NR) avec le trajet de liaison ( C1C2), sachant que r1 équivaut à la distance entre K5 et C, et en ce que K4 désigne le point de la partie côté aspiration de la ligne d'engrènement qui est la plus distante du trajet de liaison ( C1C2) entre C1 and C2, sachant que r2 équivaut à la distance entre K4 et C, et sachant que : 0.9 ≤ r 1 r 2 ≤ 0.875 × z 1 z 2 + 0.22
    Figure imgb0061
     avec z1 : nombre de dents pour le rotor secondaire (NR) et z2 : nombre de dents pour le rotor principal (HR).
  20. Paire de rotors selon l'une des revendications 1 à 19, caractérisée en ce que le rapport de longueur de rotor est conforme à : 0,85 * z 1 / z 2 + 0,67 ≤ L HR / α ≤ 1 ,26* z 1 / z 2 + 1,18 ,
    Figure imgb0062
     de préférence 0,89 ∗ (z1/z2) + 0,94 ≤ LHR/α ≤ 1,05 ∗ (z1/z2) + 1,22,
    avec z1 : nombre de dents pour le rotor secondaire (NR) et z2 : nombre de dents pour le rotor principal (HR), sachant que le rapport de longueur de rotor LHR/α indique le rapport de la longueur de rotor LHR à l'entraxe a et la longueur de rotor LHR est la distance de la face frontale de rotor principal côté aspiration à la face frontale de rotor principal opposée côté refoulement.
  21. Paire de rotors selon l'une des revendications 1 à 14, caractérisée en ce que, dans une vue en coupe frontale, le profil de dent du rotor secondaire (NR) suit, sur sa partie radialement la plus extérieure, un arc de cercle ARC1 de rayon rk1, donc plusieurs points du flanc de dent d'attaque Fv et du flanc de dent de fuite FN se situent sur l'arc de cercle de rayon rk1 autour du centre défini par l'axe C1, sachant que l'arc de cercle ARC1 inclut de préférence un angle, relativement à l'axe C1, compris entre 0,5° et 5°, de façon plus préférentielle entre 0,5° et 2,5°,
    sachant que F10 est le point le plus distant de F5 sur le flanc de dent d'attaque sur cet arc de cercle et
    sachant que la projection radiale R10 tracée entre F10 et le centre défini par l'axe C1 du rotor secondaire (NR) touche le flanc de dent d'attaque FV en au moins un point ou le croise en deux points.
  22. Bloc compresseur comprenant un carter de compresseur (15) ainsi qu'une paire de rotors selon l'une des revendications 1 à 21, sachant que la paire de rotors comprend un rotor principal (HR) et un rotor secondaire (NR) qui sont logés respectivement de façon tournante dans le carter de compresseur (15).
  23. Bloc compresseur selon la revendication 22, caractérisé en ce que la configuration en coupe frontale est réalisée de telle façon que la chambre de travail formée entre les profils de dent du rotor principal (HR) et du rotor secondaire (NR) puisse être décalée de manière sensiblement intégrale dans la fenêtre de pression.
  24. Bloc compresseur selon la revendication 22 ou 23,
    caractérisé en ce que
    un bout d'arbre du rotor principal est conduit hors du carter de compresseur et est constitué pour la liaison à un entraînement, sachant que les deux bouts d'arbre du rotor secondaire sont de préférence intégralement logés à l'intérieur du carter de compresseur.
EP15736405.0A 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse Active EP3134649B2 (fr)

Priority Applications (3)

Application Number Priority Date Filing Date Title
EP19190907.6A EP3597920B1 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse
EP18163593.9A EP3358189B9 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse
EP23198449.3A EP4273403A3 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
DE102014105882.8A DE102014105882A1 (de) 2014-04-25 2014-04-25 Rotorpaar für einen Verdichterblock einer Schraubenmaschine
PCT/EP2015/059070 WO2015162296A2 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse

Related Child Applications (5)

Application Number Title Priority Date Filing Date
EP23198449.3A Division EP4273403A3 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse
EP18163593.9A Division EP3358189B9 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse
EP18163593.9A Division-Into EP3358189B9 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse
EP19190907.6A Division EP3597920B1 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse
EP19190907.6A Division-Into EP3597920B1 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse

Publications (4)

Publication Number Publication Date
EP3134649A2 EP3134649A2 (fr) 2017-03-01
EP3134649B1 EP3134649B1 (fr) 2018-04-04
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EP3134649B2 true EP3134649B2 (fr) 2022-12-14

Family

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EP23198449.3A Pending EP4273403A3 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse
EP18163593.9A Revoked EP3358189B9 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse
EP19190907.6A Active EP3597920B1 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse
EP15736405.0A Active EP3134649B2 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse

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EP18163593.9A Revoked EP3358189B9 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse
EP19190907.6A Active EP3597920B1 (fr) 2014-04-25 2015-04-27 Paire de rotors pour bloc de compresseur d'une visseuse

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US (4) US10400769B2 (fr)
EP (4) EP4273403A3 (fr)
JP (1) JP6545787B2 (fr)
CN (1) CN106536933B (fr)
DE (2) DE102014105882A1 (fr)
ES (3) ES2668317T5 (fr)
WO (1) WO2015162296A2 (fr)

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US20220136504A1 (en) 2022-05-05
US12352266B2 (en) 2025-07-08
EP4273403A3 (fr) 2024-04-03
EP3597920B1 (fr) 2023-09-06
JP6545787B2 (ja) 2019-07-17
WO2015162296A2 (fr) 2015-10-29
EP3134649A2 (fr) 2017-03-01
US10400769B2 (en) 2019-09-03
DE102014105882A1 (de) 2015-11-12
EP3134649B1 (fr) 2018-04-04
ES2963314T3 (es) 2024-03-26
ES2668317T5 (es) 2023-04-10
EP3358189A1 (fr) 2018-08-08
EP3134649B9 (fr) 2019-02-27
ES2668317T3 (es) 2018-05-17
US20180112663A2 (en) 2018-04-26
WO2015162296A3 (fr) 2015-12-23
US11248606B2 (en) 2022-02-15
ES2967470T3 (es) 2024-04-30
US20200040894A1 (en) 2020-02-06
US20170045050A1 (en) 2017-02-16
EP4273403A2 (fr) 2023-11-08
EP3358189B9 (fr) 2024-01-03
CN106536933A (zh) 2017-03-22
CN106536933B (zh) 2019-07-12
EP3597920A3 (fr) 2021-03-24
EP3597920A2 (fr) 2020-01-22
JP2017514069A (ja) 2017-06-01
DE202015009525U1 (de) 2018-02-15
EP3358189B1 (fr) 2023-10-11
US20240175436A1 (en) 2024-05-30

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