EP0404921B1 - Rotor pair for a high-pressure screw-type compressor - Google Patents
Rotor pair for a high-pressure screw-type compressor Download PDFInfo
- Publication number
- EP0404921B1 EP0404921B1 EP90901801A EP90901801A EP0404921B1 EP 0404921 B1 EP0404921 B1 EP 0404921B1 EP 90901801 A EP90901801 A EP 90901801A EP 90901801 A EP90901801 A EP 90901801A EP 0404921 B1 EP0404921 B1 EP 0404921B1
- Authority
- EP
- European Patent Office
- Prior art keywords
- rotor
- teeth
- main rotor
- main
- auxiliary
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Expired - Lifetime
Links
Classifications
-
- 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
Definitions
- the invention relates to a pair of rotors for a high-pressure screw compressor of the type specified in the introductory part of patent claim 1.
- the tooth gaps of the main and secondary rotors form communicating V-shaped working chambers in which the working cycles include suction, compression and ejection due to the rotation of the rotors takes place and there are pressure differences on the rotors, which lead to action forces on the bearings.
- a screw compressor which comprises a total of at least three rotors, ie a central screw rotor, which is the driven main rotor, and a secondary rotor on both sides thereof.
- the main rotor is provided with concave teeth, while the two secondary rotors are provided with convex teeth.
- This type of construction differs fundamentally from the screw compressor of the type mentioned both in terms of the number of rotors and the design of the teeth on the rotors.
- the number of teeth is determined in such a way that the secondary rotor is stiff enough to prevent bending.
- Combinations of teeth with a difference in the number of secondary rotor teeth minus the number of main rotor teeth from -1 to +2 are used, which means that the load capacity of the support bearings, particularly the secondary rotor, is relatively limited due to the small center distance of the rotors.
- the area of use of known screw compressors is limited to approximately 2.5 MPa with regard to its maximum permissible operating pressure, and high-pressure applications cannot be covered.
- the invention has for its object to develop a screw compressor of the generic type in a structurally simple manner such that it allows economical work even at higher operating pressures at which high pressure screw compressors can work.
- a pair of rotors for a high-pressure screw compressor has a secondary rotor which has at least 3 teeth more than the main rotor, the difference in the number of teeth depending on the maximum operating pressure and being so great that the increased action force on the pressure-side support bearing of the associated with the increase in the number of teeth of the secondary rotor
- Auxiliary rotor is smaller than the load-bearing capacity of the largest possible support bearing, which is caused by the size of the number of secondary rotor teeth due to the increase in the center distance.
- the secondary rotor of the rotor pair according to the invention has at least 3 teeth more than the main rotor, the difference in the number of teeth depends on the maximum operating pressure and is so large that the increased action force on the pressure-side support bearing of the secondary rotor associated with the increase in the number of teeth of the secondary rotor is less than the load capacity of the largest possible support bearing, which can be used due to the increase in the center distance due to the size of the secondary rotor teeth.
- the screwing angle of the secondary rotor is reduced to 50% of the main rotor, while the center distance compared to a combination of 4 to 6 increases by 23%, so that a roller bearing as a support bearing at the secondary rotor shaft end a 40 to 50% higher load capacity (comparison basis: bearing outer diameter for the secondary rotor equals the center distance) can be used, while the action force at this point increases only by 8% at a maximum operating pressure of 3.5 MPa if the number of secondary rotor teeth is increased from 6 to 8.
- the inner column is positively influenced by the increase in the number of secondary rotor teeth, since the bearing play and thermal expansion of the rotor are in opposite directions.
- the size of the screwing angle of the main rotor is determined as a function of the peripheral speed.
- the distance between the rotors of a screw compressor which determines the size of the possible bearings, is achieved by increasing the number of secondary rotor teeth.
- the secondary rotor has at least 3 teeth more than the main rotor.
- the screwing angle of the secondary rotor is reduced to 50 percent of the main rotor.
Landscapes
- Engineering & Computer Science (AREA)
- Mechanical Engineering (AREA)
- General Engineering & Computer Science (AREA)
- Applications Or Details Of Rotary Compressors (AREA)
Abstract
Description
Die Erfindung befaßt sich mit einem Rotorpaar für einen Hochdruckschraubenverdichter der im einleitenden Teil des Patentanspruchs 1 angegeben Art. Die Zahnlücken von Haupt- und Nebenrotor bilden kommunizierende V-förmige Arbeitskammern, in denen die Arbeitszyklen umfassend das Ansaugen, Verdichten und Ausschieben infolge der Drehung der Rotoren stattfindet und dadurch Druckdifferenzen an den Rotoren vorhanden sind, die zu Aktionskräften an den Lagern führen.The invention relates to a pair of rotors for a high-pressure screw compressor of the type specified in the introductory part of patent claim 1. The tooth gaps of the main and secondary rotors form communicating V-shaped working chambers in which the working cycles include suction, compression and ejection due to the rotation of the rotors takes place and there are pressure differences on the rotors, which lead to action forces on the bearings.
In US-A-4 619 596 ist ein Schraubenverdichter der vorstehend genannten Art angegeben, welcher derart ausgelegt werden soll, daß er unter Einsatz von billigen Werkzeug einfach herzustellen sein soll, so daß sich die Dimensionierungsabstimmungen genau, wirtschaftlich und einfach vornehmen lassen. In diesem Zusammenhang ist die Wahl des Profils des Hauptrotors und der beiden Flanken des Nebenrotors von Bedeutung, wobei der Grundkreis des Profils etwa um das Zweifache größer als der Teilkreis ist. Daher ergibt sich selbst dann, wenn der Nebenrotor eine größere Zähnezahl hat, daß dieser wesentlich kleiner im Durchmesser als der Hauptrotor ist.In US-A-4 619 596 a screw compressor of the type mentioned above is specified, which should be designed such that it should be easy to manufacture using inexpensive tools, so that the dimensioning adjustments can be made accurately, economically and easily. In this context, the choice of the profile of the main rotor and the two flanks of the secondary rotor is important, the base circle of the profile being approximately twice as large as the pitch circle. Therefore, even if the secondary rotor has a larger number of teeth, it results that this is much smaller in diameter than the main rotor.
In CH-A-255 202 ist ein Schraubenverdichter angegeben, welcher insgesamt wenigstens drei Rotoren umfaßt, d. h. einen mittleren Schraubenrotor, bei dem es sich um den angetriebenen Hauptrotor handelt, und beidseitig hiervon je einen Nebenrotor. Der Hauptrotor ist hierbei mit konkaven Zähnen versehen, während die beiden Nebenrotoren mit konvexen Zähnen versehen sind. Diese Bauform unterscheidet sich sowohl hinsichtlich der Anzahl von Rotoren als auch der Auslegung der Zähne an den Rotoren grundlegend von dem Schraubenverdichter der eingangs genannten Art.In CH-A-255 202 a screw compressor is specified which comprises a total of at least three rotors, ie a central screw rotor, which is the driven main rotor, and a secondary rotor on both sides thereof. The main rotor is provided with concave teeth, while the two secondary rotors are provided with convex teeth. This type of construction differs fundamentally from the screw compressor of the type mentioned both in terms of the number of rotors and the design of the teeth on the rotors.
An intern bekannten Rotorpaaren für Schraubenverdichter ist die Zähnezahl so festgelegt, daß der Nebenrotor steif genug gegen Biegung ist.
Verwendet werden Zähnezahlkombinationen mit einer Differenz Nebenrotorzähnezahl minus Hauptretorzähnezahl von -1 bis +2, wodurch infolge des kleinen Achsabstandes der Rotoren die Tragfähigkeit der Stützlager insbesondere des Nebenrotors relativ begrenzt ist.
Dadurch wird der Einsatzbereich bekannter Schraubenverdichter hinsichtlich seines maximal zulässigen Betriebsdruckes auf etwa 2,5 MPa begrenzt, und Hochdruckeinsatzfälle können nicht abgedeckt werden.
Die Vergrößerung der Zähnezahl an Haupt- und Nebenrotor an anderen bekannten Verdichtern führt zwar zu einer unbedeutenden Verbesserung der Tragfähigkeit, beeinflußt jedoch den Liefergrad durch Verschlechterung der inneren Dichtheit ungünstig, da mit zunehmender Zähnezahl des Hauptrotors bei einem vorgegebenen Förderstrom des Verdichters die Zahnlücken, die die V-förmigen Arbeitskammern bilden, stärker verkleinert werden als die wirksamen inneren Spalte zwischen Hauptrotor- und Nebenrotorzähnen und zwischen den Rotoren und dem Gehäuse.On internally known rotor pairs for screw compressors, the number of teeth is determined in such a way that the secondary rotor is stiff enough to prevent bending.
Combinations of teeth with a difference in the number of secondary rotor teeth minus the number of main rotor teeth from -1 to +2 are used, which means that the load capacity of the support bearings, particularly the secondary rotor, is relatively limited due to the small center distance of the rotors.
As a result, the area of use of known screw compressors is limited to approximately 2.5 MPa with regard to its maximum permissible operating pressure, and high-pressure applications cannot be covered.
The increase in the number of teeth on the main and secondary rotors on other known compressors leads to an insignificant improvement in the load-bearing capacity, but adversely affects the degree of delivery due to a deterioration in the internal tightness, since with increasing number of teeth of the main rotor at a given flow rate of the compressor, the tooth gaps that the Form V-shaped working chambers, are reduced in size more than the effective inner gaps between the main rotor and secondary rotor teeth and between the rotors and the housing.
Der Erfindung liegt die Aufgabe zugrunde, einen Schraubenverdichter der gattungsgemäßen Art auf konstruktiv einfache Weise derart weiterzuentwickeln, daß er ein wirtschaftliches Arbeiten auch bei höheren Betriebsdrücken gestattet, bei welchen Hochdruckschraubenverdichter arbeiten können.The invention has for its object to develop a screw compressor of the generic type in a structurally simple manner such that it allows economical work even at higher operating pressures at which high pressure screw compressors can work.
Nach der Erfindung wird hierzu ein Rotorpaar für einen Hochdruckschraubenverdichter bereitgestellt, dessen nähere Einzelheiten im Patentanspruch 1 angegeben sind.According to the invention, a pair of rotors for a high-pressure screw compressor is provided for this purpose, the further details of which are given in patent claim 1.
Die Aufgabe der Erfindung wird dadurch gelöst, daß der Achsabstand, der die Größe der möglichen Lager begrenzt, nur durch Vergrößerung der Nebenrotorzähnezahl vergrößert wird.
Erfindungsgemäß hat ein Rotorpaar für einen Hochdruckschraubenverdichter einen Nebenrotor, der mindestens 3 Zähne mehr als der Hauptrotor aufweist, wobei die Zähnezahldifferenz vom maximalen Betriebsdruck abhängt und so groß ist, daß durch die mit der Vergrößerung der Zännezahl des Nebenrotors verbundene vergrößerte Aktionskraft auf das druckseitige Stützlager des Nebenrotors kleiner ist als die Tragfähigkeit des größtmöglichen Stützlagers, das infolge der Vergrößerung des Achsabstandes durch die Größe der Nebenrotorzähnezahl bedingt ist.
Daraus ergibt sich eine Zähnezahl des Nebenrotors, bei der die Tragfähigkeit der haupt- und nebenrotorseitigen Stützlager und Spurlager größer ist als die auftretenden gaskraftbedingten Aktionskräfte an den Lagerstellen.The object of the invention is achieved in that the center distance, which limits the size of the possible bearings, is increased only by increasing the number of secondary rotor teeth.
According to the invention, a pair of rotors for a high-pressure screw compressor has a secondary rotor which has at least 3 teeth more than the main rotor, the difference in the number of teeth depending on the maximum operating pressure and being so great that the increased action force on the pressure-side support bearing of the associated with the increase in the number of teeth of the secondary rotor Auxiliary rotor is smaller than the load-bearing capacity of the largest possible support bearing, which is caused by the size of the number of secondary rotor teeth due to the increase in the center distance.
This results in a number of teeth on the secondary rotor, in which the load capacity of the main and secondary rotor-side support bearings and thrust bearings is greater than the gas force-related action forces at the bearing points.
Weitere vorteilhafte Ausgestaltungen der Erfindung sind in den Ansprüchen 2 und 3 wiedergegeben.Further advantageous embodiments of the invention are given in claims 2 and 3.
Der Nebenrotor des Rotorpaars nach der Erfindung hat mindestens 3 Zähne mehr als der Hauptrotor, wobei die Zähnezahldifferenz vom maximalen Betriebsdruck abhängt und so groß ist, daß durch die mit der Vergrößerung der Zähnezahl des Nebenrotors verbundene vergrößerte Aktionskraft auf das druckseitige Stützlager des Nebenrotors kleiner ist als die Tragfähigkeit des größtmöglichen Stützlagers, das infolge der Vergrößerung des Achsabstandes durch die Größe der Nebenrotorzähnezahl einsetzbar ist.The secondary rotor of the rotor pair according to the invention has at least 3 teeth more than the main rotor, the difference in the number of teeth depends on the maximum operating pressure and is so large that the increased action force on the pressure-side support bearing of the secondary rotor associated with the increase in the number of teeth of the secondary rotor is less than the load capacity of the largest possible support bearing, which can be used due to the increase in the center distance due to the size of the secondary rotor teeth.
Daraus ergibt sich eine Zähnezahl des Nebenrotors, bei der die Tragfähigkeit der haupt- und nebenrotorseitigen Stützlager und Spurlager größer ist als die auftretenden gaskraftbedingten Aktionskräfte an den Lagerstellen.This results in a number of teeth on the secondary rotor, in which the load capacity of the main and secondary rotor-side support bearings and thrust bearings is greater than the gas force-related action forces at the bearing points.
Durch die erfindungsgemäße Wahl der Zähnezahlen von Hauptrotor zu Nebenrotor von 4 zu 8 reduziert sich der Verschraubungswinkel des Nebenrotors auf 50 % des Hauptrotors, während der Achsabstand gegenüber einer Kombination von 4 zu 6 um 23 % zunimmt, so daß am Nebenrotorwellenende ein Wälzlager als Stützlager mit einer um 40 bis 50 % höheren Tragfähigkeit (Vergleichsbasis: Lageraußendurchmesser für den Nebenrotor gleich Achsabstand) eingesetzt werden kann, während die Aktionskraft an dieser Stelle bei einem maximalen Betriebsdruck von 3,5 MPa nur um 8 % ansteigt, wenn die Nebenrotorzähnezahl von 6 auf 8 erhöht wird. Die inneren Spalte wird durch die Vergrößerung der Nebenrotorzähnezahl positiv beeinflußt, da Lagerspiel und Wärmedehnung des Rotors gegenläufig sind.Due to the inventive choice of the number of teeth from main rotor to secondary rotor from 4 to 8, the screwing angle of the secondary rotor is reduced to 50% of the main rotor, while the center distance compared to a combination of 4 to 6 increases by 23%, so that a roller bearing as a support bearing at the secondary rotor shaft end a 40 to 50% higher load capacity (comparison basis: bearing outer diameter for the secondary rotor equals the center distance) can be used, while the action force at this point increases only by 8% at a maximum operating pressure of 3.5 MPa if the number of secondary rotor teeth is increased from 6 to 8. The inner column is positively influenced by the increase in the number of secondary rotor teeth, since the bearing play and thermal expansion of the rotor are in opposite directions.
Ferner wird die Größe es Verschraubungswinkels des Hauptrotors in Abhängigkeit von der Umfangsgeschwindigkeit festgelegt. Dabei besteht zwischen Verschraubungswinkel φHR und Umfangsgeschwindigkeit Vu die folgende Beziehung:
wobei
K = 14 bis 18 und
Vu in Meter je Sekunde
Durch Reduzierung des Verschraubungswinkels bei kleinen Umfangsgeschwindigkeiten werden die inneren Spaltverluste insbesondere bei kleinen Schraubenverdichtern gesenkt, da der Verdichtungsvorgang schneller abläuft als bei den bisher üblichen Verschraubungswinkeln von 300 bis 325°.Furthermore, the size of the screwing angle of the main rotor is determined as a function of the peripheral speed. The following relationship exists between the screw angle φ HR and the peripheral speed V u :
in which
K = 14 to 18 and
V u in meters per second
By reducing the screwing angle at low circumferential speeds, the internal gap losses are reduced, in particular in the case of small screw compressors, since the compression process is faster than with the screwing angles of 300 to 325 ° which have been customary up to now.
Bei einem Beispiel wird der Abstand der Rotoren eines Schraubenverdichters, der die Größe der möglichen Lager bestimmt, durch die Vergrößerung der Nebenrotorzähnezahl erreicht. Der Nebenrotor hat mindestens 3 Zähne mehr als der Hauptrotor. Bei der Wahl der Zähnezahlen für Hauptrotor zu Nebenrotor von 4 zu 8 reduziert sich der Verschraubungswinkel des Nebenrotors auf 50 Prozent des Hauptrotors.In one example, the distance between the rotors of a screw compressor, which determines the size of the possible bearings, is achieved by increasing the number of secondary rotor teeth. The secondary rotor has at least 3 teeth more than the main rotor. When choosing the number of teeth for main rotor to secondary rotor from 4 to 8, the screwing angle of the secondary rotor is reduced to 50 percent of the main rotor.
Claims (3)
- Rotor pair for a high-pressure screw-type compressor, comprising a main rotor having essentially convex teeth arranged outside its reference circle and an auxiliary rotor having essentially concave teeth arranged inside its reference circle, which are screwed in opposite directions at a screwing angle based on the opposite end faces of main rotor to auxiliary rotor, the rotors being mounted coaxially in support or step bearings at the axial distance of their rotary axes, which absorb the action forces occurring in the radial and axial direction during the operating cycle comprising intake, compression and discharge, characterised in that the auxiliary rotor has at least three teeth more than the main rotor to increase the axial distance between the rotors, such that the load-bearing capacity of the pressure-side support bearing of the auxiliary rotor is increased more than the increase in the action force acting thereon due to the greater tooth number of the auxiliary rotor during the operating cycles.
- Rotor pair for a high-pressure screw-type compressor according to claim 1, characterized in that the screwing angle of the main rotor measured in degrees is 14 to 18 times as large as the numerical value of the peripheral speed of the main rotor in metres per second.
- Rotor pair for a high-pressure screw-type compressor according to claim 1 or 2, characterised in that the main rotor has four teeth and the auxiliary rotor has eight teeth for a maximum operating pressure of 3.5 MPa.
Applications Claiming Priority (2)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| DD325008 | 1989-01-17 | ||
| DD325088A DD301062A7 (en) | 1989-01-17 | 1989-01-17 | Rotor pair for high pressure screw compressor |
Publications (2)
| Publication Number | Publication Date |
|---|---|
| EP0404921A1 EP0404921A1 (en) | 1991-01-02 |
| EP0404921B1 true EP0404921B1 (en) | 1993-11-10 |
Family
ID=5606651
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| EP90901801A Expired - Lifetime EP0404921B1 (en) | 1989-01-17 | 1990-01-17 | Rotor pair for a high-pressure screw-type compressor |
Country Status (6)
| Country | Link |
|---|---|
| US (1) | US5096399A (en) |
| EP (1) | EP0404921B1 (en) |
| JP (1) | JP2812797B2 (en) |
| DD (1) | DD301062A7 (en) |
| DE (1) | DE59003410D1 (en) |
| WO (1) | WO1990008261A1 (en) |
Families Citing this family (4)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| DE102006021704B4 (en) * | 2006-05-10 | 2018-01-04 | Gea Refrigeration Germany Gmbh | Screw compressor for large power outputs |
| DE102006035782B4 (en) * | 2006-08-01 | 2018-10-25 | Gea Refrigeration Germany Gmbh | Screw compressor for extremely high operating pressures |
| DE102006035783A1 (en) * | 2006-08-01 | 2008-02-07 | Grasso Gmbh Refrigeration Technology | screw compressors |
| GB0921968D0 (en) * | 2009-12-17 | 2010-02-03 | Epicam Ltd | A rotary deviceand method of designingand makinga rotary device |
Family Cites Families (6)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| CH255202A (en) * | 1944-06-29 | 1948-06-15 | Ljungstroms Angturbin Ab | Rotary machine. |
| DE3111834A1 (en) * | 1981-03-26 | 1982-10-21 | Balcke-Dürr AG, 4030 Ratingen | Screw compressor |
| DE3140108A1 (en) * | 1981-10-09 | 1983-04-28 | Technika Beteiligungsgesellschaft mbH, 4800 Bielefeld | TURN PISTON COMPRESSORS |
| JPS60116920A (en) * | 1983-11-30 | 1985-06-24 | Hitachi Ltd | Thrust bearing holding in rotary fluid machine |
| GB8413619D0 (en) * | 1984-05-29 | 1984-07-04 | Compair Ind Ltd | Screw rotor machines |
| SE453318B (en) * | 1987-02-18 | 1988-01-25 | Svenska Rotor Maskiner Ab | ROTOR MACHINE WITH AN AXIAL POWER BALANCING DEVICE |
-
1989
- 1989-01-17 DD DD325088A patent/DD301062A7/en unknown
-
1990
- 1990-01-17 JP JP2502330A patent/JP2812797B2/en not_active Expired - Fee Related
- 1990-01-17 DE DE90901801T patent/DE59003410D1/en not_active Expired - Lifetime
- 1990-01-17 EP EP90901801A patent/EP0404921B1/en not_active Expired - Lifetime
- 1990-01-17 WO PCT/EP1990/000093 patent/WO1990008261A1/en active IP Right Grant
- 1990-07-18 US US07/553,802 patent/US5096399A/en not_active Expired - Lifetime
Also Published As
| Publication number | Publication date |
|---|---|
| DD301062A7 (en) | 1992-10-01 |
| JPH03503307A (en) | 1991-07-25 |
| US5096399A (en) | 1992-03-17 |
| EP0404921A1 (en) | 1991-01-02 |
| WO1990008261A1 (en) | 1990-07-26 |
| DE59003410D1 (en) | 1993-12-16 |
| JP2812797B2 (en) | 1998-10-22 |
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