EP1682779B1

EP1682779B1 - Radial compressor impeller

Info

Publication number: EP1682779B1
Application number: EP04768299.2A
Authority: EP
Inventors: Andrew John Vine; Keith Robert Pullen
Original assignee: Dynamic Boosting Systems Ltd
Current assignee: Dynamic Boosting Systems Ltd
Priority date: 2003-09-09
Filing date: 2004-09-02
Publication date: 2016-11-09
Anticipated expiration: 2024-09-02
Also published as: GB0321088D0; WO2005024242A1; EP1682779A1

Description

The invention relates to a compressor, in particular a compressor of the radial or centrifugal type for a turbo machine.
The basic principle of radial compressors is well known: the compressor comprises a rotor driven by a turbine or other machine having a plurality of generally radial blades which divert axially flowing inlet gas such as air at the centre to provide a pressure rise at the circumference exit. Often diffusers in the form of tangential vanes are provided to slow down airflow at the exit and hence convert the kinetic energy of the airflow to a pressure rise.
The type of compressor used is dependant upon factors such as gas volume flow and one type of compressor particularly useful for flow rates in the intermediate region (1 to 50 litres per second) is described in GB2366333 . According to this document a multi-stage compressor comprises a plurality of concentrically nested rotors in a correspondingly nested stator configuration. Each stator stage has diffuser vanes and a flow passage to the axial air inlet in the next rotor stage. In one embodiment described in GB2366333 the vane profiles are effectively laminar or plate like occupying a minimal volume of the compressor space and this is termed here a "full entry" compressor. A particularly effective embodiment described in GB2366333 , however, relies on a wedge shaped blade occupying a substantial fraction of the compressor volume and this is termed here a "partial entry" compressor.
A problem with full entry turbomachines (standard radial geometry) is that they must operate at a particular speed for a given flow rate and pressure rise - the science behind this can be quantified using the concept of specific speed. If the flow rate is relatively low, the shaft speed must be high in order to maintain the correct physical dimensions. This creates a problem in that once speeds get over about 20,000 rpm, it is not easy to find a drive system. The drives do exist but are expensive - ie a low speed motor with a gearbox or a high speed motor (inverter driven). Once speeds get above 100,000 rpm then it is very difficult to find an appropriate drive. Using partial entry is a way of reducing flow rate without increasing shaft speed. Taking a radial compressor with a flow rate of 1m³/s and pressure ratio of 3:1 it would operate at about 30,000 rpm. To obtain ¼ of the flow, the geometry is scaled down by 2 in linear terms and operates at twice the speed. The inlet area goes down by 4 hence ¼ the flow and the pressure ratio is the same since the rotor tip speed is the same. However the machine operates at 60,000 rpm. An alternative is to take the original machine and block off ¾ of the passages. The gas in the remaining passages suffers no effect as long as the inlet is adjusted (no blunt surfaces).
The main advantage of partial entry compressors is being able to operate at a much reduced shaft speed in comparison to conventional radial compressors. For example - a compressor for 10 m³/s can operate at 60,000 rpm as opposed to 600,000 rpm. It can be used as a single stage - generally the pressure ratio is limited to 1.6:1 but in most cases, a multistage device is required to achieve higher pressure ratios. The pressure ratio is the multiple of the pressure ratio for each stage and the number of stages. 5 stages each of pressure ratio 1.6:1 can achieve 10.48:1 (neglecting interstage pressure losses).
However there are various problems with partial entry systems. Firstly the partial entry compressor provides an interupted flow of gas to the diffuser at the exit. This is because the flow passages only occupy a fraction of the available area, the rest of which is contained within enclosed islands comprising the partial entry blades. When the flow leaves the rotor, it is in the form of a number of rotating discrete jets of number equal to the number of rotor passages. Such a flow will enter the diffuser whose purpose is to recover as much kinetic energy as possible and convert this to useable static pressure rise. However, since the diffuser will experience a pulsating flow from the jets, the efficiency of diffusion is quite poor.
A further problem is windage loss caused by movement of rotor parts adjacent to generally stationary gas causing the gas to move, drawing power from the rotor which is not useful. The windage problem is severe for partial entry compressors, increasing approximately with the fifth power of diameter of the rotor. Although the nested configuration solves some of the problem because rotors are adjacent to moving gas from the rotor from which they are nested , the use of partial entry rotors still means that the ratio windage losses to useful work goes up by a factor of 4.
US patent number 3,369,737 relates to a radial flow machine and, more particularly, to a radial outflow compressor that pumps to high compression ratio in a single stage.
The invention is set out in the claims. In particular because of the configuration of the partial entry blades a large amount of the forward speed is imparted to the gas flowing through the flow passages allowing the respective flows to re-meet hence minimising the wakes or pulsation effect found in known arrangements.
Embodiments of the invention will now be described, by way of example, with reference to the drawings, of which:

Fig. 1 shows a cross-sectional view of a radial compressor according to the present invention; and
Fig. 2 shows a cross-sectional view of a radial compressor having an alternative vane configuration.

In overview the invention relates to a radial compressor having swept forward partial entry blades, that is, the blades are curved forwardly in the direction of rotation. In particular the forward sweep is turned extensively towards the tangential direction in the direction of rotor rotation such that the resultant exit flow from the exaggerated forward swept flow passage has a tangential velocity greater than the velocity of the compressor blade tips. As a result it is found that the flow passages effectively reform to provide a full flow passage around the periphery of the motor significantly reducing the effect of the jets at the exit and hence increasing the efficiency of the diffuser. Furthermore an improved pressure ratio is achieved as a result of which the diameter of the compressor can be reduced and/or the compressor can be used in a single stage rather than requiring a nested configuration. This reduction in dimension allows smaller seals and less leakage is encountered giving a higher overall efficiency.
Fig. 1 shows a rotor 10 having a plurality of partial entry blades 12. The rotor 10 is driven by a shaft 13 such that air inducted at an axial inlet passage 14 is driven generally outwardly in the passages 16 between adjacent blades 12 to a circumferential exit at 18 when the compressor rotor is rotating in the direction shown by arrow A. The exit air is diffused by a plurality of generally linear, tangentially extending diffusers 20 which are angled in the direction of rotation and are wedge shaped, tapering inwardly to a point adjacent the circumferential exit 18 of the compressor rotor 10. The diffuser passage wall is preferably close in the radial direction to the rotor exit guiding the flow in an almost tangential direction, maintaining the correct flow angle at the rotor exit and hence maintaining the required pressure ratio. The increased pressure air exits the diffuser to the load or to another stage as appropriate.
The rotor blade 12 can be solid or hollow and includes a concave forward face 22 in the direction of flow A and an increased curvature concave rear face 24 forming generally a D shape profile pointing away from the direction of flow. The blade 22 occupies a significant proportion of the volume of the rotor space as a result, a "dead space" being defined between the front and rear faces. The forward face is angled generally tangentially and in the direction of flow at the radially innermost inlet region 26 and curves through approximately 180 degrees to extend generally tangentially once again at the radially outer most exit region 28. The opposing rear face 24 of an adjacent blade 12 is profiled to provide a curved flow passage 16 therebetween which exits generally tangentially and is of generally constant width. The specific profile of the blades/volumes of the blades depends on the gas being compressed and the rotor speed and can be optimised for each case as will be apparent to the skilled reader. The exit blade angle is preferably between 20 degrees and 90 degrees (tangential) to a radius of the rotor, as long as sufficient forward speed is provided to allow the flows in the passages of the compressor to re-converge, minimising the pulsation effect. It will be appreciated that the rotor is also profiled in the axial direction but this can be in an entirely conventional manner which will be apparent to the skilled person and so is not described here.
The significance of the exaggerated forward sweep of the invention can be understood with reference to a general discussion of conventional profiling of compressor blades. In a backswept design, where the rotors are curved away from the direction of flow, the pressure ratio decreases with volume flow rate because the absolute rotor exit velocity is decreased. However in a forward swept design, the pressure ratio increases with volumetric flow rate, therefore there is no need to increase the static pressure in the rotor, as this will only decrease the pressure ratio. Keeping a similar pressure at rotor inlet and exit is beneficial for 2 reasons. Firstly an unfavorable pressure gradient can cause the flow to separate, especially if there is a great deal of turning as in this design. Secondly the flow at the rotor exit normally has a tendency to leak past the rotor shroud on the outside of the rotor because the static pressure at the rotor exit is higher than the inlet pressure. If there is no static pressure rise in the rotor, then this problem is eliminated.
The pressure rise of a turbo compressor is a direct function of the change in tangential velocity of the gas in a rotor such that the greater the change in velocity, the greater the pressure rise. In particular, the work input to the gas depends on the change in tangential velocity multiplied by the blade speed, and pressure ratio is a direct function of work input. As a result the exit velocity increases with the tip speed of the blades and hence the diameter of the rotor. With the addition of the exaggerated forward sweep of the present invention the gas is forced to jet forward faster than the blade speed, the tangential velocity is greater still at the exit and hence so is the pressure increase. Since the flow leaves the rotor at a velocity greater than the tip speed of the rotor at gas exit, the blade speed (product of rotor radius and shaft speed) can be reduced and the pressure ratio can be achieved with a lower than normal diameter, in comparison with radial, moderately forward swept or back swept passages. Since the diameter of the rotor is lower, the windage losses are significantly reduced as they are approximately proportionate of the fifth power of diameter of the rotor. For the partial entry machine, it is worth suffering the penalty of low efficiency due to forward sweep since the reduction in windage losses more than makes up for this.
By way of example, a rotor with moderately forward swept blades requires a diameter of 90mm at a speed of 60,000 rpm in order to achieve a pressure ratio of 1.6:1 in air. This same pressure ratio can be achieved according to the invention with a rotor diameter of only 70mm at the same speed and a decrease in windage losses to about 30% of the original value.
It will be appreciated that a range of possible blade profiles and diffuser profiles can be adopted and it will be seen that an alternative diffuser vane profile is shown in Figure 2 in which the vanes 20 are generally curved forwardly in the direction of rotation A.
Because the diameter of the rotor 10 can be reduced, the size of the seals are also reduced and because there is little static pressure rise in the rotor (due to the forward sweep), leakage losses from rotor tip to inlet are minimal. The types of seal available will be familiar to the skilled reader and are described in more detail, in GB2366333 .
It will be appreciated that individual elements from the embodiments described above can be interchanged or juxtaposed as appropriate. The compressor described can be used in a single or multi-stage arrangement and any multi-stage arrangement can be nested or a more conventional axial system. The compressor can be driven by any appropriate machine such as a turbine or electrical machine and can be used in any appropriate implementation.

Claims

A centrifugal compressor comprising a plurality of forward swept partial entry rotor blades (12) each having a forward face (22) in the direction of rotation, and further comprising a diffuser having a plurality of diffuser passage walls (20); characterized in that the rotor blades (22) each include an exit portion (28) extending substantially tangentially to the rotor circumference and the diffuser passage walls (20) each extend substantially tangentially to the rotor circumference and are positioned sufficiently close to the exit portion (28) in a radial direction to guide the air flow in a generally tangential direction.
A centrifugal compressor according to claim 1, in which the plurality of rotor blades (12) define therebetween forward swept flow passages (16), the flow passages (16) being sufficiently forward swept to allow convergence of flow streams from individual passages in the vicinity of the passage exit (18).
A centrifugal compressor according to any preceding claim, in which each rotor blade (12) has a forward face (22) in the direction of rotation and a rear face (24) defining an area therebetween.
A centrifugal compressor according to claim 3, in which the forward face (22) of the rotor blade (12) is concave in the direction of rotation and the rear face (24) of the rotor blade (12) has increased curvature and is convex pointing away from the direction of rotation.
A centrifugal compressor according to claim 4, in which the forward face (22) is angled generally tangentially and in the direction of rotation at its radially innermost region (26) and curves through approximately 180 degrees to extend generally tangentially once again at its radially outermost region (28).
A centrifugal compressor according to any of claims 3 to 5 in which the forward (22) and rear (24) faces of respective adjacent rotor blades (12) define a flow passage (16) therebetween.
A centrifugal compressor according to claim 6 in which the flow passage (16) between rotor blades (12) is curved in the direction of forward sweep.
A centrifugal compressor according to claim 6 or claim 7 in which the flow passage (16) between adjacent rotor blades (12) is of substantially constant width.
A centrifugal compressor according to any of the preceding claims in which the diffuser passage walls (20) are generally forward swept in the direction of rotation.
A centrifugal compressor according to any of claims 6 to 9, in which the maximum tangential width of each rotor blade is significantly greater than the width of each flow passage at the point on the passage which is adjacent the widest part of the blade.
A centrifugal compressor according to any of claims 6 to 9, in which the maximum tangential width of each blade is at least three times the width of each flow passage at the point on the passage which is adjacent the widest part of the blade.
A centrifugal compressor according to any of the preceding claims comprising a plurality of rotors (10) in a multi-stage configuration.