GB1598914A

GB1598914A - Centrifugal compressors

Info

Publication number: GB1598914A
Application number: GB3819777A
Authority: GB
Original assignee: GOSLING M; Compair Industrial Ltd
Current assignee: GOSLING M; Compair Industrial Ltd
Priority date: 1978-05-09
Filing date: 1978-05-09
Publication date: 1981-09-23

Description

(54) IMPROVEMENTS IN CENTRIFUGAL COMPRESSORS (71) We, COMPAIR INDUSTRIAL LIMITED, a British Company, of P.O. Box 7, Broomwade Works, High Wycombe, Buckinghamshire, HP13 5SF, and MARTIN CHARLES GOSLING, a British Subject, of 465 West Wycombe Road, High Wycombe, Buckinghamshire, do hereby declare the invention for which we pray that a patent may be granted to us and the method by which it is to be performed, to be particularly described in and by the following statement: The invention relates to compressors and more specifically to centrifugal compressors.

The invention provides a centrifugal compressor comprising two stages, each stage having its own impeller rotatably mounted in a housing and having an air inlet and an air outlet, the outlet of the first stage being connected directly or indirectly to the inlet of the second stage, in which the two impellers are mounted for rotation on a single shaft, and the impeller of the first stage is larger than the impeller of the second stage, the relative sizes of the impellers being such that the work loading of the first stage is substantially greater than the work loading of the second stage.

Preferably the relative sizes of the impellers are such that the ratio of the work loading (total enthalpy rise) of the first stage to the work loading of the second stage is approximately 11:5.

Preferably the shaft includes an integrally forrned gear and preferably said gear is in mesh with a parallel-shaft gear train. Preferably drive is applied to one of the gears of said parallel-shaft train via an epicyclic gear train.

The gears of the parallel-shaft train may be arranged between the housings of the first and second stages. In this case, the impellers are preferably mounted facing in opposed directions at opposite ends of the single shaft with the integrally formed gear between the two impellers.

The arrangement may be such that the aerodynamic thrust of the two impellers is at least partially balanced.

A preferred embodiment of the invention will now be described, by way of example, with reference to the accompanying drawing which is a section through a two-stage compressor according to the invention.

The classical thermodynamic analysis of a two stage compressor with complete intercooling suggests that for minimum overall work input at a given output load, each stage should be absorbing an equal share of the work. However, for small capacity turbo machines the very small scale of the second stage introduces practical problems associated with small running clearances, etc. For practical reasons these critical clearances become proportionately greater with the reduced overall scale and the second stage performance suffers. To reduce the effects of this it is possible to make the first stage absorb more than half of the total work input. In the machine described below there is optimised, the relative contribution of each stage for a single shaft 2 stage machine (i.e. a machine in which the impeller of each stage rotates at the same speed).All types of turbo machines have an opitmum operating point defined typically by a specific speed parameter. By adjusting the pressure ratio and hence enthalpy rise of each stage it is possible for both stages to operate at their optimum specific speed.

The drawing shows a two-stage compressor, each stage comprising an impeller 10, 11 rotatably mounted within a housing 12, 13. The two impellers 10, 11 are mounted at opposite ends of a single shaft 15, each impeller being attached to the shaft end in known manner.

The shaft 15 is supported in bearings 16, 17 one said bearing being located in each housing 12, 13. The bearings are of well-known type for this application, and shaft seals, 19, 20, also of known type fit around portions of the shaft 15 between the bearings 16, 17 and the respective impellers 10, 11.

The single shaft 15 includes an integrally formed gear 22 at its central part, through which gear drive is applied to the shaft. The gear 22 is in mesh with a gear 23, the shaft of which is parallel to the shaft 15 and is supported between the housings 12, 13 by bearings 25, 26, one said bearing being located in each housing. The gear 23 is in turn in mesh with a further parallel shaft gear 27 also mounted between the housings 12, 13 in bearings 28, 29. The gear 27 has an internally splined connection to a drive shaft 30 and the shaft 30 is driven through an epicyclic gear train (not shown but located at 34) by a prime mover in known fashion.

The two housings 12, 13 are maintained in spaced relationship by a spacer ring 31 to which the housings are bolted.

Ambient air enters the first compressor stage axially at 32 and after compression by the impeller 10 passes radially out of the first stage via a diffuser 33 formed in the housing, and thence to a collector 35, in the form of a toroidal duct of varying cross-sectional area.

The partially compressed air is then taken from collector 35 via an inter cooler (not shown) and connecting ducts (not shown) and is fed axially into the second compression stage at 37. This air is then compressed by the impeller 11 of the second stage and the compressed air passes radially out of this stage via diffuser 38 to a collector 39. The compressed air from the collector 39 is then supplied for use, usually via an after cooler (not shown).

It will be seen from the drawing that the two impeller stages are of different physical size. In fact, the relative size of the two impeller stages is so selected that, when driven at the same speed on the single shaft 15, both stages are operating at optimum conditions.

In the typical example shown, the ratio of the work input to the stages is 11:5 for an overall pressure ratio of approximately 8:1.

The work input ratio is selected to achieve two things. First, this ratio allows both stages to operate at approximately their optimum specific speed.

Second, it is known that the efficiency of an impeller stage falls as its design capacity decreases due to scale effects, this effect is much stronger with the second stage than the first. Thus, by having a relatively heavily loaded first stage and a lightly loaded second stage, it is possible to partially offset this effect and the overall efficiency of the machine is kept as high as possible.

As with all "back-to-back" two stage single shaft machines, the aerodynamic thrust of one impeller is at least partially balanced by that of the other.

WHAT WE CLAIM IS : - 1. A centrifugal compressor comprising two stages, each stage having its own impeller rotatably mounted in a housing and having an air inlet and an air outlet, the outlet of the first stage being connected directly or indirectly to the inlet of the second stage, in which the two impellers are mounted for rotation on a single shaft, and the impeller of the first stage is larger than the impeller of the second stage, the relative sizes of the impellers being such that the work loading of the first stage is substantially greater than the work loading of the second stage.

2. A compressor as claimed in claim 1 in which the relative sizes of the impellers are such that the ratio of the work loading of the first stage to the work loading of the second stage is approximately 11: 5.

3. A compressor as claimed in claim 1 or claim 2 in which the shaft includes an integrally formed gear.

4. A compressor as claimed in claim 3 in which said gear is in mesh with a parallel shaft gear train.

5. A compressor as claimed in claim 4 in which drive is applied to one of the gears in said train via an epicyclic gear train.

6. A compressor as claimed in claim 4 or claim 5 in which the parallel shaft gear train is arranged between the housings of the first and second stages.

7. A compressor as claimed in claim 6 in which the impellers are mounted facing in opposed directions at opposite ends of the single shaft with the integrally formed gear between the two impellers.

8. A compressor as claimed in any one of the preceding claims in which the aerodynamic thrust of the two impellers is at least partially balanced.

9. A centrifugal compressor substantially as hereinbefore described with reference to and as shown in the accompanying drawings.

**WARNING** end of DESC field may overlap start of CLMS **.

Claims

**WARNING** start of CLMS field may overlap end of DESC **. in each housing. The gear 23 is in turn in mesh with a further parallel shaft gear 27 also mounted between the housings 12, 13 in bearings 28, 29. The gear 27 has an internally splined connection to a drive shaft 30 and the shaft 30 is driven through an epicyclic gear train (not shown but located at 34) by a prime mover in known fashion. The two housings 12, 13 are maintained in spaced relationship by a spacer ring 31 to which the housings are bolted. Ambient air enters the first compressor stage axially at 32 and after compression by the impeller 10 passes radially out of the first stage via a diffuser 33 formed in the housing, and thence to a collector 35, in the form of a toroidal duct of varying cross-sectional area. The partially compressed air is then taken from collector 35 via an inter cooler (not shown) and connecting ducts (not shown) and is fed axially into the second compression stage at 37. This air is then compressed by the impeller 11 of the second stage and the compressed air passes radially out of this stage via diffuser 38 to a collector 39. The compressed air from the collector 39 is then supplied for use, usually via an after cooler (not shown). It will be seen from the drawing that the two impeller stages are of different physical size. In fact, the relative size of the two impeller stages is so selected that, when driven at the same speed on the single shaft 15, both stages are operating at optimum conditions. In the typical example shown, the ratio of the work input to the stages is 11:5 for an overall pressure ratio of approximately 8:1. The work input ratio is selected to achieve two things. First, this ratio allows both stages to operate at approximately their optimum specific speed. Second, it is known that the efficiency of an impeller stage falls as its design capacity decreases due to scale effects, this effect is much stronger with the second stage than the first. Thus, by having a relatively heavily loaded first stage and a lightly loaded second stage, it is possible to partially offset this effect and the overall efficiency of the machine is kept as high as possible. As with all "back-to-back" two stage single shaft machines, the aerodynamic thrust of one impeller is at least partially balanced by that of the other. WHAT WE CLAIM IS : -

1. A centrifugal compressor comprising two stages, each stage having its own impeller rotatably mounted in a housing and having an air inlet and an air outlet, the outlet of the first stage being connected directly or indirectly to the inlet of the second stage, in which the two impellers are mounted for rotation on a single shaft, and the impeller of the first stage is larger than the impeller of the second stage, the relative sizes of the impellers being such that the work loading of the first stage is substantially greater than the work loading of the second stage.