EP0592803A1 - Gear driven multi-shaft compressor with return stages and radial-expander - Google Patents

Abstract

The invention relates to a geared multi-shaft turbo compressor with impellers (8, 8a) connected in series in terms of flow, which are fastened to two or more pinion shafts (6) arranged in parallel to one another, which are connected directly via a central wheel (5) or indirectly via pinion shafts on the circumference of the Central wheel (5) are driven. In the case of the low-pressure stages (first or first and second pinion shaft) following the second or third pinion shaft (6), there are several impellers (8, 8a) in succession with the interposition of a disc diffuser (9) and a return ring (10) on at least one pinion shaft end (6) arranged. By reversing the direction of flow, i.e. H. Entry of the gas on the high pressure side and exit of the gas on the low pressure side as well as reversing the direction of rotation results in a radial expander (turbine) with the same basic design. <IMAGE>

Description

The invention relates to a multi-stage geared multi-shaft turbocompressor with flow-connected impellers which are fastened to two or more pinion shafts arranged in parallel to one another, which are driven directly via a central wheel or indirectly via pinion shafts on the circumference of the central wheel.

The external drive can be an electric motor, a steam or gas turbine, etc. in a known manner.

In the case of indirect drive, the power can be transferred to the compressor impellers via the pinion shaft of the drive via the central wheel via the pinion shaft of the compressor impeller or the central wheel via idler gears via the pinion shaft of the compressor impeller.

In the multi-shaft turbo compressors according to the prior art, for. B. DE-PS 974 418, on each pinion shaft on one or both sides of the pinion, a compressor impeller is arranged flying. Connecting pipes are located between the stages.

The gas enters the impeller axially via the intake housing and is decelerated in the volute casing. With increasing compression and thus a decreasing intake volume flow of the compressor stages, the impellers in the outer diameter become smaller and smaller in order to maintain optimal volume flow numbers and the speeds of the pinion shafts to maintain the peripheral speed of the impellers required for the respective stage compression ratio. This leads to the maximum diameter of the central wheel given by its maximum peripheral speed ever smaller pinion diameters and number of pinion teeth.

When the limit number of teeth is reached, intermediate gears have to be activated in order to be able to increase the speed further. This leads to further mechanical friction losses in the gearbox bearings and gear teeth.

At high speeds, rotor-dynamic problems with regard to vibration stability, etc. can also occur.

The solution shown in DE-OS 25 15 628, in which a pair of impellers on each pinion shaft is only arranged on one side on the side of the transmission facing away from the drive, offers rotor dynamics rather deteriorations, especially here due to the between the transmission housing and impeller arranged radial intake manifold creates a large center of gravity distance of the overhanging rotor part.

An intermediate cooler is normally arranged between the individual compressor stages, which cools the gas back down to the initial temperature of the compression. As a result, the end temperatures of the individual compressor stages are correspondingly low, corresponding to the temperature increase of the stage.

If the process also requires a high final temperature, the final stage must run at a correspondingly high peripheral speed in order to achieve the required final temperature. This increases the pinion shaft speed even more, which means that problems mentioned above are further aggravated.

One way out of avoiding the high peripheral speed would be to increase the number of stage pressures, e.g. B. by steeper impeller exit angles, but this deteriorates the characteristic curve behavior with respect to the surge limit.

Another option would be to connect two stages in series with a connecting pipe without an intercooler. In addition to the additional construction work for a second pinion shaft end and two complete spiral stages, there are additional flow losses due to the double energy conversion of pressure and speed energy, additional leakage losses at the exit of the pinion shaft from the spiral housing and mechanical friction losses.

The object of the invention is to provide a multi-shaft turbo-compressor which avoids the disadvantages of the prior art and which is characterized in that in multi-shaft turbo-compressors, in particular with high total pressure ratios, a perfect mechanical behavior with high overall efficiency and low construction costs is realized.

The object is achieved by the features listed in the main claim. Advantageous embodiments of the device are described in subclaims 2-16. In the claims 17-20, the features of a radial expander (turbine) are described, which arises by reversing the direction of rotation and flow of the geared multi-shaft turbo compressor.

Claim 21 describes the features according to which both variants are arranged in a common machine.

The solution to the object is achieved in the multi-shaft turbo compressor according to the invention in that in the low pressure stages (first or first and second pinion shaft) following stages from the second or third pinion shaft, several impellers are arranged one behind the other with the interposition of a disc diffuser and a feedback ring on at least one pinion shaft end .

The low-pressure stages can be designed as conventional individual stages, which run in the usual way with a high peripheral speed and great swallowing capacity and thus already greatly reduce the volume flow.

The suction to the first impeller of the high-pressure stage group, which is formed from one or more recirculation stages and an end spiral stage, takes place via an axial inlet connection.

The disc diffuser connected to the impeller can be designed without blades or with diffuser guide vanes.

The direct transfer of the outlet flow from the disc diffuser via the return ring into the subsequent stage and the arrangement of the last stage of each stage group directly next to the gearbox housing result in a compact design that minimizes the center of gravity of the impellers from the support bearing arranged in the gearbox housing.

The direct transfer of the outlet flow to the subsequent stages of a stage group also avoids pressure losses due to a double pressure conversion (deceleration to pipeline speed and subsequent acceleration to the impeller inlet speed of the subsequent stage).

By dividing the specific compression work to be performed by an impeller with a high peripheral speed and speed into two or more stages, the speed can be reduced considerably. In spite of the larger wave overhang, more favorable conditions can be achieved in terms of rotor dynamics.

The following advantages in terms of flow technology result:
While maintaining the volume flow numbers, the impeller diameter increases when the speed is reduced, while maintaining the impeller diameter, the volume flow numbers increase. Both effects have a positive influence on the flow efficiency given the small impeller diameters used especially in the high-pressure section and the small volume flow numbers that are often present.

If an arrangement of one high-pressure stage group on each of the two shaft ends of a pinion shaft runs in opposite directions in the direction of flow, there is extensive compensation between the oppositely directed axial thrusts caused by flow forces. If, on the other hand, there is only one high-pressure stage group on one shaft end, a relief piston is arranged on the other shaft end, in particular with regard to changing operating states, if for this relief piston there is no more space on the shaft end with the step group according to the invention, which has a larger shaft overhang than a single step, taking into account the position of the critical speeds.

This relief piston is particularly well suited to changing operating pressures of the compressor if the gas generating the axial thrust is directed from the wheel chamber behind the last stage of the high-pressure stage group arranged on the same pinion shaft to the rear of the relief piston and the gas drawn in by the high-pressure stage group the outer end of the relief piston is passed.

For process-related reasons, it may be necessary to attach inlet or outlet connections to the return rings of the uncooled stage groups if the inlet or outlet pressure specified by the process is between the inlet and outlet pressure of a high-pressure stage group.

The impellers can be connected to one another via spur gears, suitably a Hirth serration. This enables a horizontal, undivided design of the housing rings as in the conventional single stage.

The spur toothing consists of radial grooves that are machined into the end faces of the impellers. These interlock, are radially centered and transmit the torque.

The toothed components are held together axially by a central expansion screw that is screwed into the pinion shaft. The spur gear elements can also be manufactured separately and attached to the impellers.

In order to reduce the number of axially releasable shaft connections, because of the radial displacements of the rotor components that are possible within the scope of their manufacturing tolerances, it may be necessary to firmly connect the impellers of a group of steps to each other, e.g. B. by a shrink connection and only with a common spur gear on the pinion shaft. This requires a horizontal parting line on the housing ring arranged between the impellers of a step group. The outer housing, including the suction-side and pressure-side covers, can remain undivided.

The spur gear can be arranged in the connected hub of the impeller group so that it is located approximately in the center of gravity of the impellers.

According to a further embodiment of the invention, the inner housing is designed with a horizontal parting joint and surrounded by a horizontally undivided outer housing. In this case, the entire rotor can be installed in the gearbox without disassembly after balancing. In contrast to the pot design with single-shaft compressors, no horizontally undivided cover on the gear unit side can seal off the casing.

Gas leakage through the remaining horizontal parting line to the outside atmosphere at high gas pressures is avoided by lowering the intermediate pressure in the housing chamber between the inner housing and the gear housing via relief lines.

In the case of a multi-stage arrangement of the impellers, it is expedient because of the thermal expansion of the rotor to design the impellers with a cover disk. This is also possible at low peripheral speeds of the high-pressure stage group compared to the single-stage version with high peripheral speeds.

If, despite the reduction in the pinion shaft speed due to the longer overhanging shaft end of a high-pressure stage group, the critical speeds become too low, according to the invention the first impeller of the stage group is used to reduce the rotor mass and shift the center of mass with a smaller outer diameter than the impellers of the subsequent stages and / or, if necessary, without a cover plate. Other variants consist of designing one or more impellers from a material with a density below that of steel, for example titanium or aluminum alloys.

In an advantageous embodiment of the invention, rotor dynamic problems are solved, in particular in the case of widely overhanging rotors in the high pressure range, by the use of active magnetic bearings which hold the rotor in position by sensors and have controllable damping.

In addition to the radial magnetic bearings for receiving the remaining axial thrust, the known pressure combs on the gear pinions or separate axial magnetic bearings can be used.

Particularly in the case of the relatively long shaft overhang of the high-pressure stage groups according to the invention, subsynchronous shaft vibrations which can occur at high gas pressures must be avoided. In order to reliably avoid the swirl currents that cause them in the labyrinths of the impeller and shaft seals, the housing walls of the wheel chambers are provided with swirl breaking grooves in this case, which take the swirl out of the leakage current before it enters the labyrinth seals. For a residual swirl to be expected, the labyrinth seals on the leakage current inlet side are equipped with swirl breaking ribs arranged perpendicular to the circumferential direction. In addition, sealing gas is conducted without swirl or with counter-swirl from the radially outer area of the wheel chambers into the labyrinth, which prevents rotating leakage currents from entering the labyrinth seal from the wheel chamber.

To achieve a wide working range with a high partial load efficiency, axial guide vanes and secondary guide vanes with adjustable diffuser vanes are used in compressors.

In the step groups considered here, it proves to be expedient in terms of construction and flow technology to equip the first step of a step group with an axial guide wheel and the last step with an adjustable guide wheel in front of the end spiral.

By reversing the direction of flow of the geared multiwave turbo machine designed as a geared multiwave turbo compressor, i. H. Entry of the gas on the high pressure side and exit of the gas on the low pressure side when the direction of rotation is reversed, the geared multi-shaft turbo machine works as a radial expander with the same basic design. Compared to the conventional design, the step arrangement according to the invention in the high-pressure part achieves a constant or even greater gradient per pinion shaft end with good vibration stability.

Here the outlet spiral of the compressor becomes the inlet spiral of the radial expander, the non-bladed or bladed disc diffuser becomes the inlet guide wheel, the intake manifold of the step group becomes the outlet diffuser. The return ring can be carried out with or without blades.

Only the profiling of the blade grids of the rotor, guide and return blades is adapted to the reverse flow direction.

With high pressure ratios, ie here with large enthalpy drops, the gas volume flows at the beginning of the expansion are still very small and require small impeller diameters to achieve optimal volume flow rates. The speeds of the pinion shafts must be correspondingly high in order to maintain the peripheral speed required for the enthalpy gradient in question. This would lead to the same gearbox problems as with the compressors in a conventional step arrangement.

The new design also offers advantages when combining compressors and radial expanders in a common gear housing. The construction effort can be reduced by combining high-pressure stage groups of compressors and radial expanders on a common pinion shaft. The number of compressor or radial expander stages of a pinion shaft can be varied and optimized in order to adjust the optimal speeds for a given step pressure ratio and enthalpy gradient.

Fig. 1

eine stirnseitige Ansicht eines mehrstufigen Getriebe-Mehrwellenturbokompressors nach dem Stand der Technik mit drei Ritzelwellen,

Fig. 2

einen Schnitt A-A durch die untere horizontale Teilfuge eines Turbokompressors nach Fig. 1,

Fig. 3

einen Schnitt B-B durch die obere horizontale Teilfuge eines Turbokompressors nach Fig. 1, dessen Niederdruckteil nach Fig. 2 ausgeführt ist,

Fig. 4

einen vertikalen Schnitt durch ein Ritzelwellenende eines Getriebe-Mehrwellenturbokompressors nach dem Stand der Technik, gemäß Fig. 1,

Fig. 5

einen Schnitt A-A durch die untere horizontale Teilfuge eines erfindungsgemäßen Turbokompressors mit einer konventionellen Niederdruck-Welle und einer neuartigen Hochdruck-Welle mit den Stufen III bis VI,

Fig. 6

einen Schnitt A-A durch die untere horizontale Teilfuge eines erfindungsgemäßen Turbokompressors mit neuartigen Stufen IV und V,

Fig. 7

einen Schnitt B-B durch die obere Teilfuge eines erfindungsgemäßen Turbokompressors, dessen Niederdruck-Teil nach Fig. 2 ausgeführt ist, mit je zwei neuartigen Stufen V, VI, VII, VIII an den Ritzelwellenenden,

Fig. 8

einen Schnitt B-B durch die obere horizontale Teilfuge eines erfindungsgemäßen Turbokompressors mit zwei neuartigen Stufen und einem Entlastungskolben am anderen Ritzelwellenende,

Fig. 9

einen horizontalen Schnitt durch ein Ritzelwellenende ähnlich Fig. 7 mit in das Laufrad integriertem Entlastungskolben,

Fig.10

einen horizontalen Schnitt durch ein Ritzelwellenende mit einer ersten Stufe ohne Deckscheibe,

Fig.11

einen vertikalen Schnitt durch ein Ritzelwellenende mit zwei auf die Welle geschrumpften Laufrädern,

Fig.12

einen horizontalen Schnitt entspr. Fig. 10 mit zusätzlichen Einspeisungskanälen,

Fig.13

einen horizontalen Schnitt entspr. Fig. 10 mit zusätzlichen Entnahmekanälen,

Fig.14

einen Schnitt durch die obere horizontale Teilfuge eines erfindungsgemäßen Turbokompressors mit aktiver Magnetlagerung,

Fig.15

einen Schnitt durch die obere horizontale Teilfuge eines erfindungsgemäßen Turbokompressors mit radialer aktiver Magnetlagerung und axialen Druckkämmen am Ritzel,

Fig.16

einen horizontalen Schnitt durch ein Ritzelwellenende ähnlich Fig. 7 mit zwei Laufrädern mit Deckscheibe und nur einer Stirnverzahnung,

Fig.17

einen Schnitt durch ein Ritzelwellenende eines erfindungsgemäßen Turbokompressors mit einer Hirthverzahnung im Schwerpunkt der Stufen,

Fig.18

einen Schnitt durch ein Wellenende eines erfindungsgemäßen Turbokompressors mit Drallbrechern und Sperrgaseinleitung,

Fig.19

einen vertikalen Schnitt durch ein Wellenende eines erfindungsgemäßen Turbokompressors mit axialem Vorleitrad vor der ersten Stufe und radialem Nachleitrad in der letzten Stufe,

Fig.20

einen Schnitt A-A durch die untere horizontale Teilfuge eines erfindungsgemäßen Radialexpanders mit einer Ritzelwelle mit den Hochdruckstufen I bis IV und einer Ritzelwelle mit konventionellen Niederdruckstufen V und VI,

Fig.21

einen horizontalen Schnitt durch ein Ritzelwellenende eines Radialexpanders.

Fig.22

einen Schnitt A-A durch die horizontale Teilfuge einer Getriebe-Mehrwellenturbomaschine mit einer Ritzelwelle mit den Hochdruckstufen II und III eines Turbokompressors sowie den Hochdruckstufen I und II eines Radialexpanders und einer Ritzelwelle mit der Niederdruckstufe I des Turbokompressors und III des Radialexpanders.

Embodiments of the invention are illustrated with the aid of schematic drawings. The individual shows:

Fig. 1

an end view of a multi-stage gear multi-shaft turbo compressor according to the prior art with three pinion shafts,

Fig. 2

2 shows a section AA through the lower horizontal parting line of a turbo compressor according to FIG. 1,

Fig. 3

2 shows a section BB through the upper horizontal parting joint of a turbo compressor according to FIG. 1, the low pressure part of which is designed according to FIG. 2,

Fig. 4

2 shows a vertical section through a pinion shaft end of a geared multi-shaft turbo compressor according to the prior art, according to FIG. 1,

Fig. 5

3 shows a section AA through the lower horizontal parting joint of a turbo compressor according to the invention with a conventional low-pressure shaft and a novel high-pressure shaft with stages III to VI,

Fig. 6

3 shows a section AA through the lower horizontal parting line of a turbo compressor according to the invention with novel stages IV and V,

Fig. 7

FIG. 2 shows a section BB through the upper part of a turbocompressor according to the invention, the low-pressure part of which is designed according to FIG. 2, each with two novel stages V, VI, VII, VIII at the pinion shaft ends,

Fig. 8

1 shows a section BB through the upper horizontal parting joint of a turbo compressor according to the invention with two novel stages and a relief piston at the other end of the pinion shaft,

Fig. 9

7 shows a horizontal section through a pinion shaft end similar to FIG. 7 with a relief piston integrated in the impeller,

Fig. 10

a horizontal section through a pinion shaft end with a first stage without a cover disk,

Fig. 11

a vertical section through a pinion shaft end with two impellers shrunk onto the shaft,

Fig. 12

10 shows a horizontal section corresponding to FIG. 10 with additional feed channels,

Fig. 13

10 shows a horizontal section corresponding to FIG. 10 with additional removal channels,

Fig. 14

2 shows a section through the upper horizontal parting line of a turbo compressor according to the invention with active magnetic bearing,

Fig. 15

3 shows a section through the upper horizontal parting joint of a turbo compressor according to the invention with radial active magnetic bearing and axial pressure combs on the pinion,

Fig. 16

a horizontal section through a pinion shaft end 7 with two impellers with a cover disk and only one spur toothing,

Fig. 17

2 shows a section through a pinion shaft end of a turbo compressor according to the invention with serration in the center of gravity of the stages,

Fig. 18

2 shows a section through a shaft end of a turbo compressor according to the invention with swirl crushers and sealing gas inlet,

Fig. 19

4 shows a vertical section through a shaft end of a turbo compressor according to the invention with an axial guide wheel in front of the first stage and a radial guide wheel in the last stage,

Fig. 20

3 shows a section AA through the lower horizontal parting joint of a radial expander according to the invention with a pinion shaft with high pressure stages I to IV and a pinion shaft with conventional low pressure stages V and VI,

Fig. 21

a horizontal section through a pinion shaft end of a radial expander.

Fig. 22

a section AA through the horizontal joint of a geared multi-shaft turbomachine with a pinion shaft with the high pressure stages II and III of a turbo compressor and the high pressure stages I and II of a radial expander and a pinion shaft with the low pressure stage I of the turbo compressor and III of the radial expander.

Fig. 1 shows the front view of a known turbo compressor. Three compressor stages with a spiral housing (2) are attached to a gearbox housing (1) and are driven by a central drive shaft (3) or a pinion shaft (4) arranged on the circumference of the central wheel.

Fig. 2 shows a section through the lower part of such a turbo compressor.

The gas enters the impeller (8) via the intake housing (7). The gas flow is delayed in the volute casing (2).

The impellers of stages I to IV are dimensioned smaller and smaller due to the increasing compression in order to maintain optimal volume flow rates in the outer diameter.

In Fig. 3, a section through the upper horizontal parting line of a turbo compressor according to Fig. 1, structural details such as gear (5, 6), impellers (8a), housing (1), etc. can be seen. The low pressure part is designed according to FIG. 2.

FIG. 4 illustrates in a vertical section through a pinion shaft end (6) structural features of the multi-shaft turbo compressor of the prior art according to FIG. 1.

5 shows the schematic structure of a turbo compressor according to the invention.

The turbo compressor with the volute casing (2) and the intake manifold (7) is equipped with a conventional low-pressure shaft (6) with stages I and II and a high-pressure shaft (6) according to the invention with stages III to VI.

Two compressor impellers (8a) are arranged on the same pinion shaft end in the same flow direction on the high-pressure shaft (6). Disc diffusers (9) and return rings (10) are interposed.

Fig. 6 is a section through the lower horizontal parting of a turbo compressor according to the invention with the high pressure stages IV and V according to the invention, wherein two impellers (8a) are arranged on the pinion shaft (6). Disc diffusers (9) and return rings (10) are also interposed here.

From Fig. 7, a section through the upper horizontal parting line of a turbo compressor according to the invention, one can see design details of two high pressure stages (V, VI and VII, VIII) at the pinion shaft ends (6). The low pressure part is in this turbo compressor in a conventional manner. Fig. 2 executed. The first impeller (8a) of the high-pressure stage groups has a reduced outer diameter. The impeller is fastened with the help of the well-known Hirth toothing, a spur toothing (11) with a central fastening screw (12).

The axially opposing arrangement of the two stage groups compensates for the axial thrusts generated by each stage group due to the pressure differences before and after the impeller.

8, a compressor with two high-pressure stages at a pinion shaft end (6), a relief piston (15) is arranged at the opposite pinion shaft end within a pressure-resistant housing (13), which serves to compensate for axial thrusts.

In this example, compressed gas is supplied from the wheel chamber (27) via the line (24a) to the inner chamber (28a) on the relief piston, while the outer chamber (28) via the relief line (24) to the suction port (7) of the first stage of the Step group is lowered in the pressure level.

Fig. 9, a horizontal section through a pinion shaft end (6), shows the design with two compressor impellers with a cover plate (8a), both impellers (8a) having the same outside diameter. The inner housing (17) is undivided and a relief piston (15) is integrated in the second impeller (8a).

Fig. 10 shows a horizontal section of a pinion shaft end (6) with an undivided inner housing of another design (17a). The first impeller (8) has no cover disk and has a smaller outer diameter than the next stage with cover disk (8a).

11 shows the pinion shaft end (6) of a turbo compressor according to the invention with two impellers (8a) shrunk onto the pinion shaft (6) with a shaft sleeve (29) arranged in between. The compressor inner housing (18) is divided horizontally and screwed with its lower part to the gear housing. The inner housing upper part (18a) is after Insert the pinion shaft (6) with the inner housing lower part (18b) screwed.

The undivided outer housing (19) is then pushed over and axially screwed to the gear housing middle (25a) and upper part (25), whereby an additional housing chamber (26) is formed, which can be relieved of pressure via the relief line (24).

According to the horizontal section according to FIG. 12, a turbocompressor according to FIG. 10 additionally has gas feed channels (20) between the compressor stages, which end in the suction-side housing cover (30).

In Fig. 13, a sectional view corresponding to Fig. 10, additional gas extraction channels (21) can be seen, which are shown between the two compressor stages shown and end in the suction-side housing cover (30).

Fig. 14, a section through the upper horizontal parting of a geared multi-shaft turbo compressor according to the invention with the impellers (8a), is intended to indicate the radial (22) and the axial magnetic bearing (23), which compensate for dynamic problems by the magnetic bearings over Sensors hold the rotor (6) in the desired position.

Fig. 15, a section through the upper horizontal parting of a geared multi-shaft turbo compressor according to the invention with the impellers (8a), shows radial magnetic bearings (22). The remaining axial thrust is absorbed in a conventional manner by pressure combs (39) via the central wheel (5) from the axial pressure bearing of the central wheel shaft (3).

Fig. 16, a horizontal section through a pinion shaft end (6), shows the design with two compressor impellers with a cover plate (8a), both impellers (8a) having the same outside diameter. Both impellers are firmly connected to each other, here the impeller (8a) with cover plate is shown shrunk on the extended hub of the impeller (8b). As a result, only a serration (11) is required, but the inner housing (18) must be horizontally divided (18a, 18b) for installation. A relief piston (15) is integrated in the second impeller (8b).

17 shows structural details of an impeller attachment (8a, 8b). The second impeller (8b) with an extended hub of the high-pressure stage group encloses with its extended hub the pinion shaft end (6), in the end face of which a Hirth toothing is milled. In the elongated hub, a ring (11a) with counter-Hirth teeth is inserted on a projection (42) for manufacturing reasons. The first impeller (8a) is firmly connected (shrunk, soldered, welded) to the second impeller (8b) via a centering (43).

Both impellers (8a, 8b) are attached to the pinion shaft end (6) together with the central fastening screw (12).

18 and FIGS. 18a-18d show details of swirl breakers and the introduction of sealing gas.

The letters A, B, C and D shown in FIG. 18 denote the enlarged sections in FIGS. 18a-18d.

In the impeller chamber (27) of the first and second impeller (8a, 8b), on the cover plate side and in the impeller chamber of the second impeller (8b), on the wheel disc side, radial swirl breaking grooves (35) are incorporated, which prevent leakage from the outer surface of the impeller to the labyrinth seals (36) The impellers (8a, 8b), shaft (6) and the relief piston (15) should break. Swirl breaking ribs (37) are arranged in the labyrinth seals (36) on the gas inlet side perpendicular to the circumferential direction and are intended to destroy swirl components of the flow velocity that have entered the labyrinth seal (36).

As a result of the pressure difference between the radially outer region of the wheel chambers (27) and the suction opening of the impellers (8a, 8b), a sealing gas flow is passed through the bores (38) into the labyrinth seal (36) of the impellers (8a, 8b) in order to avoid any Prevent any swirling flow from entering the impeller chamber adjacent to the labyrinth seal (36). The same applies to the relief piston (15).

The labyrinth seal (36) on the intermediate sleeve (40) between the stages is supplied with sealing gas from the wheel chamber of the subsequent stage via the bores (38).

FIG. 19 shows a guide vane (31) with adjusting device (34) in front of the first stage of a compressor and a guide vane (32) with adjusting device (32a) after the second stage.

20 shows the schematic structure of a radial expander according to the invention through the lower horizontal parting line.

The radial expander is equipped with a high-pressure shaft (6) according to the invention with high-pressure stages I to IV and a conventional low-pressure shaft (6) with stages V and VI. Two expander impellers (8a) are arranged on the same pinion shaft end (6) in the same flow direction on the high-pressure shaft (6).

From the inlet housing (2a), which is designed as a spiral housing, and the stator (33a) arranged in the disc annulus (9a), the gas enters the impeller (8a) and then via the return ring (10a) into the second stage, from there into the outlet cone diffuser (7a) of the radial expander.

21 shows in a horizontal section a pinion shaft end (6) of a radial expander with an undivided inner housing (17a). At the inlet of the impellers, inlet guide wheels (33a) are arranged in the disk annulus (9a). The return ring (10a) is designed here without a blade and serves for deflection and as a radial diffuser after the first impeller (8a).

22 shows the combination of a geared multi-shaft turbomachine with a turbocompressor according to the invention (left side of the picture) with a radial expander (right side of the picture), the turbocompressor compressing a medium other than that in the radial expander. In the high pressure area of the compression of the turbo compressor (stage groups II and III) and in the expansion in the radial expander (stage groups I and II), the different volume flows are small and allow the pinion shaft speed to be the same. By the arrangement on a common pinion shaft (6) reduces the construction costs of the combined geared multi-shaft turbomachine and the axial thrusts are largely compensated for.

In the low-pressure range of compression of the compressor (stage I) and expansion of the radial expander (stage III), the volume flows are of the same order of magnitude, so that the arrangement of the stages in question on a common pinion shaft (6) also offers advantages.

List of reference numbers:

1

Gear housing

2nd

Spiral casing for compressors

2a

Turbine inlet casing

3rd

Central shaft drive shaft

4th

Pinion shaft, on the circumference of the central wheel, for external drive

5

Central wheel

6

Pinion shaft with compressor and / or radial expander stages

7

Intake manifold for compressor stages

7a

Outlet diffuser for radial expander stages

8th

Impeller without a cover disk

8a

Impeller with a cover disc

8b

Impeller with an extended hub

9

Disc diffuser

9a

Disc ring space with radial expanders

10th

Return ring of the compressor

10a

Return ring of the radial expander

11

Spur toothing (according to Hirth)

11a

Ring with Hirth serration

12th

central fastening screw for spur gearing

13

pressure-resistant housing for relief pistons

14

Shaft seal

14a

Horizontally split shaft seal upper part

14b

Horizontally divided shaft seal lower part

15

Relief piston

16

Seal of 15

17th

undivided inner casing

17a

undivided inner housing (different version)

18th

split inner housing

18a

Top of 18

18b

Lower part of 18

19th

undivided outer housing

20th

Feed channel

21

Extraction channel

22

radial magnetic bearing

23

axial magnetic bearing

24th

Discharge line

24a

Line to the relief piston

25th

Gearbox upper part from 1

25a

Gearbox middle part from 1

25b

Gearbox lower part from 1

26

Housing chamber

27

Wheel chamber

27a

Housing wall of the wheel chamber

28

Outer chamber on the relief piston

28a

Inner chamber on the relief piston

29

Shaft bushing

30th

suction-side housing cover

30a

Insert for the housing cover on the suction side

31

Idler

32

Guide wheel

32a

Adjustment device guide pulley

33

Guide vanes

33a

Radial expander inlet guide wheel

34

Adjuster wheel

35

Radial swirl breaking grooves

36

Labyrinth seal

36a

Labyrinth rings

37

Swirl breaking ribs

38

Sealing gas supply

39

Pressure comb

40

Intermediate socket

41

poetry

42

Protrusion of the impeller hub

43

Centering in the impeller hub

I, II, III, IV, V, VI, VI, VII, VIII

Sequence of pressure levels in the direction of flow.

Claims (21)

Multi-stage geared multi-shaft turbo compressor with impellers connected in series in terms of flow, two or more compressor impellers being fastened on one or more pinion shafts arranged in parallel to one another, which are driven directly via a central wheel or indirectly via pinion shafts on the circumference of the central wheel
characterized,
that in the low-pressure stages (first or first and second pinion shaft) following stages from the second or third pinion shaft (6) as a high-pressure stage group, several impellers (8, 8a) in succession with the interposition of a disc diffuser (9) and a feedback ring (10) on at least one pinion shaft end ( 6) are arranged. Multi-stage gear multi-shaft turbo compressor according to claim 1,
characterized,
that on one side of a pinion shaft (6) a high pressure stage group and on the other side alone a relief piston (15) is arranged. Multi-stage gear multi-shaft turbo compressor according to claims 1 to 2,
characterized,
that on one or more return rings (10) connecting piece for feeding (20) or removal (21) of gas are arranged to increase or decrease the gas flow conveyed. Multi-stage geared multi-shaft turbo compressor according to claim 1 and one or more of the following claims,
characterized,
that impellers (8, 8a) of high pressure stages are connected to one another by spur gears (11) and central bolts (12). Multi-stage geared multi-shaft turbo compressor according to claim 1 and one or more of the following claims,
characterized,
that the inner housing (18) is designed with a horizontal parting joint and the horizontally undivided outer housing (19) surrounds the divided inner housing (18) with the rotor, that the gear housing (1) with the divided inner housing (18) forms an additional housing chamber (26) , and that a relief line (24) is connected to the housing chamber (26). Multi-stage geared multi-shaft turbo compressor according to claim 1 and one or more of the following claims,
characterized,
that the impellers (8a) of a step group are designed with a cover plate. Multi-stage geared multi-shaft turbo compressor according to claim 1 and one or more of the following claims,
characterized,
that the impeller (8) of the first stage of a stage group is designed without a cover plate. Multi-stage geared multi-shaft turbo compressor according to claim 1 and one or more of the following claims,
characterized,
that the impeller (8, 8a) of the first stage of a stage group has a smaller outer diameter (D) than the stage connected via the return ring (10). Multi-stage geared multi-shaft turbo compressor according to claim 1 and one or more of the following claims,
characterized,
that one or more impellers (8 or 8a) of a step group are made of material of lower density than steel. Multi-stage geared multi-shaft turbo compressor according to claim 1 and one or more of the following claims,
characterized,
that one or more pinion shafts (6) with high pressure stages are mounted in magnetic bearings (22, 23). Multi-stage geared multi-shaft turbo compressor according to claim 1 and one or more of the following claims,
characterized,
that one or more pinion shafts (6) are mounted in radial magnetic bearings (22) and the relevant pinion shafts (6) and the central wheel (5) have axial pressure combs (39). Multi-stage geared multi-shaft turbo compressor according to claim 1 and one or more of the following claims,
characterized,
that two or more impellers (8 or 8a) of a step group with a common spur gear (11) are attached to the pinion shaft end (6). Multi-stage gear multi-shaft turbo compressor according to claim 12,
characterized,
that the common spur toothing (11) is arranged in the region of the center of gravity of the impellers (8 and 8a). Multi-stage geared multi-shaft turbo compressor according to claim 1 and one or more of the following claims,
characterized,
that housing walls of the wheel chambers (27a) are provided with radial swirl breaking grooves (35). Multi-stage geared multi-shaft turbo compressor according to claim 1 and one or more of the following claims,
characterized,
that in labyrinth seals (36) in the central area of the labyrinths sealing gas feeds (38) open and in the edge areas of the leakage current entry side of the labyrinth swirl breaking ribs (37) are present perpendicular to the circumferential direction. Multi-stage geared multi-shaft turbo compressor according to claim 1 and one or more of the following claims,
characterized,
that in front of the first stage of a stage group an adjustable guide wheel (31) and in the last stage of the stage group an adjustable guide wheel (32) are arranged. Multi-stage gear multi-shaft turbo compressor according to claims 1 to 15,
characterized,
that this is designed as a radial expander (turbine) by reversing the direction of flow, ie entry of the gas on the high pressure side (2a) and exit of the gas on the low pressure side (7a). Multi-stage gear multi-shaft turbo compressor according to claim 17,
characterized,
that the outlet spiral (2) of the high pressure stages of the compressor as an inlet spiral (2a) one Radial expander, the disc diffuser (9) of the high-pressure stages of the compressor is used as the inlet annular space (9a) of a radial expander,
that the return ring of the high-pressure stages (10) of the compressor is designed as a return ring (10a) of the radial expander and the intake manifold (7) of the high-pressure stages of the compressor as an outlet diffuser (7a) of the radial expander. Multi-stage gear multi-shaft turbo compressor according to claims 17 and 18,
characterized,
that the diffuser guide vanes (33) of the high-pressure stages of the compressor are designed as an inlet guide wheel (33a) of the radial expander. Multi-stage gear multi-shaft turbo compressor according to claim 19,
characterized,
that the return ring (10a) of the radial expander is not bladed. Multi-stage geared multi-shaft turbo compressor according to claim 1 and one or more of the following claims,
characterized,
that high pressure stage groups of turbo compressors and radial expanders for different media are arranged on a common pinion shaft (6).

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