EP0440902B1 - Transmission and centrifugal compressor - Google Patents

Description

The invention relates to a multi-stage turbo compressor driven by a gear, a so-called gear turbo compressor with the features of the preamble of claim 1.

In a known two-stage geared turbo compressor (FR-A-885 488), the impellers of the two compressor stages are arranged on the ends of a shaft. The shaft is driven by a drive gear via an intermediate gear designed as a double gear. Another known turbocompressor (Dubbel Taschenbuch für den Maschinenbau 14th edition, 1981, pages 920-921) has four compressor stages, the impellers of which are arranged in pairs on two pinion shafts and are driven by a central drive gear of a single-stage transmission.

sind zur Erzielung hoher Wirkungsgrade sieben Verdichterstufen und bei Druckverhältnissen P_A / P_E von größer als 80 mindestens acht Verdichterstufen erforderlich. Der Volumenstrom ändert sich von Stufe zu Stufe erheblich. Zur Erhaltung optimaler Verhältnisse sind entsprechende Drehzahlerhöhungen nach jeweils zwei Stufen erforderlich. Die für einen optimalen Verdichter zu realisierenden Drehzahlen der Antriebswelle und der Ritzelwelle der letzten Stufe sind so verschieden, daß die notwendige Übersetzung von größer als 1 : 24 nicht mehr in ein übliches Getriebe eingebaut werden kann.The achievable efficiency of a compressor stage of a turbo compressor depends on the inlet volume flow, speed and delivery head. With pressure ratios between outlet pressure P _A and inlet pressure P _E of over ${\text{P}}_{\text{A}}{\text{/ P}}_{\text{E}}\text{= 60}$

seven compressor stages are required to achieve high levels of efficiency and at least eight compressor stages are required for pressure ratios P _A / P _E of greater than 80. The volume flow changes considerably from level to level. To maintain optimal conditions, corresponding speed increases are required after every two stages. The speeds of the drive shaft and the pinion shaft of the last stage to be realized for an optimal compressor are so different that the necessary ratio of greater than 1:24 can no longer be installed in a conventional transmission.

ausgelegt worden. Hierbei ist ein einstufiges Getriebe mit bis zu drei Ritzelwellen ausgerüstet. Jede Ritzelwelle treibt zwei direkt aufgesetzte Laufräder an, die zwei Verdichterstufen bilden. Bei einem sechsstufigen Verdichter und einem durchschnittlichen Druckverhältnis von 1,8 pro Stufe läßt sich ein Druckverhältnis von ${\text{P}}_{\text{A}}{\text{/ P}}_{\text{E}}\text{= 34}$

erzielen. Hochbelastete Stufen mit einem durchschnittlichen Druckverhältnis von 2,0 pro Stufe würden ein Gesamtdruckverhältnis von ${\text{P}}_{\text{A}}{\text{/ P}}_{\text{E}}\text{= 64}$

ergeben. Ein solcher Verdichter wäre nur wenig regelbar, und der Gesamtwirkungsgrad wäre wegen hoher Machzahlen sehr ungünstig.Five-stage and six-stage geared turbo compressors are up to now for maximum pressure ratios ${\text{P}}_{\text{A}}{\text{/ P}}_{\text{E}}\text{= 40}$

been interpreted. Here, a single-stage gear is equipped with up to three pinion shafts. Each pinion shaft drives two directly attached impellers that form two compressor stages. With a six-stage compressor and an average pressure ratio of 1.8 per stage, a pressure ratio can be achieved from ${\text{P}}_{\text{A}}{\text{/ P}}_{\text{E}}\text{= 34}$

achieve. Highly stressed stages with an average pressure ratio of 2.0 per stage would have a total pressure ratio of ${\text{P}}_{\text{A}}{\text{/ P}}_{\text{E}}\text{= 64}$

surrender. Such a compressor would be difficult to regulate and the overall efficiency would be very unfavorable due to the high Mach numbers.

z. B. würde in der siebten Stufe eine Drehzahl von 35 000 1/min erfordern. Nur 28 000 1/min können mit herkömmlichen Getrieben für Turboverdichter realisiert werden. Deshalb ist der Gesamtwirkungsgrad nicht optimal. Es besteht zwar die Möglichkeit, die vierte Ritzelwelle mit einer achten Verdichterstufe zu besetzen. Aber auch dann bleibt die grundsätzliche Problematik der Drehzahlbegrenzung mit den geschilderten Konsequenzen bestehen.Pressure ratios higher than 60 can only be achieved by adding a seventh or eighth compressor stage if the indispensable properties such as controllability and high overall efficiency are to be retained. However, the highest levels of efficiency have so far failed due to the fact that the greatest feasible transmission ratio in the known transmissions for turbocompressors is approximately 1:20. For high outputs and drives with electric motors with four poles, this means a maximum speed of the last pinion shaft of 30,000 rpm. If this maximum speed is taken as a basis, all turbo compressors with volume flows below 60,000 m³ / h and pressure ratios above 60 are no longer able to achieve uniformly optimal efficiencies, since the step delivery heights for the first steps would have to be considerably higher than for the last. The first stages thus have the highest energy density with only moderate efficiencies due to the inevitably high Mach numbers. A seven-stage turbocompressor with a volume flow of 55,000 m³ / h and a pressure ratio of ${\text{P}}_{\text{A}}{\text{/ P}}_{\text{E}}\text{= 64}$

e.g. B. would require a speed of 35,000 rpm in the seventh stage. Only 28,000 rpm can be achieved with conventional gearboxes for turbo compressors. Therefore the overall efficiency is not optimal. It is possible to fill the fourth pinion shaft with an eighth compressor stage. But even then the fundamental problem of speed limitation with the consequences outlined remains.

The invention has for its object to build a transmission turbocompressor for high pressure ratios with good efficiencies and sufficient controllability.

This object is achieved in a generic turbo compressor according to the invention by the characterizing features of claim 1.

The intermediate gear, designed as a double gear with different gear modules, creates a gearbox for turbo compressors which, in addition to moderate speeds for the low compressor stages, also realizes very high speeds for the higher compressor stages without the number of teeth of the central drive gear being assigned to the number of teeth assigned to the higher compressor stages Pinion depends. The low speeds with high tooth forces require wide toothing and a large tooth module, the high speeds with correspondingly small tooth forces require narrower wheels with a fine tooth module. The intermediate gearwheel with the different tooth modules also allows the combination of two independent gearboxes for turbocompressors in one housing and allows the installation of ratios up to 40. This makes it possible to achieve optimum efficiency even with geared turbo compressors for high pressure ratios. The overall efficiency improvement is more than 5%, which means considerable savings in energy costs.

An embodiment of the invention is shown in the drawing and is explained in more detail below. The drawing shows a gearbox for a multi-stage turbocompressor.

The transmission shown serves to drive a multi-stage turbocompressor and contains a central drive gear 2 connected to a drive shaft 1. The drive gear 2 meshes with two pinions, which are each arranged on a pinion shaft 3, 4. The impeller of a first stage and the impeller of a second stage of the turbocompressor are arranged on the ends of the pinion shaft 3. The pinion shaft 4 carries at the ends the impeller of a third and a fourth stage, which cannot be seen in the drawing. All impellers revolve in a spiral housing.

The drive gear 2 is followed by an intermediate gear 5, which is designed as a double gear and consists of a first gear 51 and a second gear 52, which are connected to one another and arranged on a common shaft. The first gear 51 meshes with the drive gear 2, while the second gear 52 engages with pinions of further pinion shafts 6, 7, 8. The pinion shafts 6, 7, 8 each have an impeller on both shaft ends, which rotates in a spiral housing. The pinion shaft 6 is assigned the fifth and sixth compressor stages, the pinion shaft 7 the seventh and eighth compressor stages and the pinion shaft 8 the ninth and tenth compressor stages.

The second gear 52 of the intermediate gear 5 has a larger diameter than the first gear 51, which is in engagement with the drive gear 2, so that the pressure combs of the pinion shafts 6, 7, 8 can overlap via the second gear 52. Furthermore, the first gear 51 has a smaller number of teeth and a larger tooth module in comparison to the second gear 52.

The interposition of the intermediate gearwheel 5 described makes it possible, owing to the diversity of the tooth modules, to make the first four stages completely independent of the subsequent stages in terms of transmission technology. The necessary translation of the pinion shaft 4 is approximately 10.5. If the number of teeth with Z2 = 25 were selected for the pinion of the pinion shaft 4, the drive gearwheel 2 would have ZO = 263 teeth, which would result in a diameter of DO = 1575 mm and a tooth engagement speed of 122 m / s with a tooth module of 6. It will, as usual in the known gearboxes for turbo compressors, determine the diameter of the housing spiral of the first and second compressor stage, the center distance of the pinion shafts 3 and 4, which in this example will be 1800 m, with a diameter of the drive gear 2 of 1596 mm and a number of teeth of ZO = 266.

The first gear 51 of the intermediate gear 5, which is in engagement with the drive gear 2, has approximately Z51 = 132 teeth and a diameter of 790 mm at a speed of 3000 1 / min.

With an eight-stage compressor, the pinion shaft 7 has a speed of 34730 1 / min. Assuming a number of teeth of Z4 = 21, the second gear would have 52 Z52 = 243 teeth. With a tooth module of 4, this results in a diameter of 972 mm and a tooth engagement speed of 150 m / s. The total moment of inertia remains acceptable due to the moderate diameter and allows the drive motor upstream of the gearbox to start the drive shaft without problems.

Claims (2)

1. Multistage turbocompressor with an integrated gear which displays a central input gear wheel (2) and an intermediate gear wheel (5), which is constructed as double gear wheel with a first gear wheel (51) and a second gear wheel (52) and the first gear wheel (51) of which stands in engagement with the central input gear wheel (2), wherein the impeller wheels of the compressor stages are arranged on the ends of pinion shafts (3, 4, 6, 7, 8) and the pinion shafts stand in engagement with the second gear wheel (52) of the intermediate gear wheel (5), characterised thereby, that the first gear wheel (51) of the intermediate gear wheel (5) displays a greater tooth modulus than the second gear hweel (52) and that the first pinion shafts (3, 4) are driven directly by the central input gear wheel (2) and only the further pinion shafts are driven by the second gear wheel (52) of the intermediate gear wheel (5).
2. Multistage turbo-compressor according to claim 1, characterised thereby, that the first gear wheel (51) of the intermediate gear wheel (5) displays a smaller diameter than the second gear wheel (52).

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