Field of the Invention

[0001] This invention relates to a multiple-entry fluid turbine in which the fluid passes through a single bladed or double-sided bladed disc more than once between the inlet and the exhaust.

Background to the Invention

[0002] In an early form of steam turbine, known as the Elektra turbine, the steam flow expands radially inwards and outwards through the same rotor via a series of partial admissions. This type of turbine enables multiple expansion stages to be achieved using a single rotor, permitting a combination of low steam flow, high expansion and low shaft speed. Turbines of this type are conventionally of the velocity compounded type, in which the bulk of the pressure drop occurs in the inlet nozzle and, predominantly, velocity drops through each partial admission (or stage). The pressure remains essentially constant at the turbine disc, minimising leakage of the working fluid, but the flow angle at each stage reduces, leading to a progressive reduction in stage efficiency due to worsening flow incidence.

[0003] In GB2463660A1 , we disclosed a re-entry turbine in which each of the expansion stages has a dynamically variable nozzle in which the flow areas can be varied according to the fluid flow rate through the turbine, the object being to maintain the fluid velocities of each stage at, or close to, the design point condition. This means that the turbine efficiency at the design point can be retained across a significant load range.

[0004] Typically multistage, full admission, axial turbines are employed in large power plants, for example fossil-fuelled and nuclear steam power plants generating hundreds of MW, and these can achieve high efficiencies. However for smaller-scale power generation, and in particular for waste-heat recovery, full admission axial-flow turbines can be inappropriate because of cost and complexity and because of the general inverse relationship between turbine size and shaft speed. Full admission radial turbines are favoured over axial tur- bines at small scale, but full admission radial turbines also suffer from the inverse size/speed relationship. High shaft speed is generally undesirable as it necessitates the use of a high-speed alternator or a gearbox. The re-entry type of turbine is being considered for small-scale waste heat recovery applications, as is disclosed, for example, in Weiss, Andreas P, et al, “Numerical and Experimental Investigation of a Velocity Compounded Radial Re-Entry Turbine for Small-Scale Waste Heat Recovery”, Energies 2022, 15, 45, 30 December 2021.

[0005] An advantage of re-entry turbines of this type is that they can operate at much lower rotational speeds than equivalent duty full admission axial or radial turbines, but studies to date suggest that efficiencies of around 45% can be achieved with velocity compounded radial re-entry turbines. The present invention seeks to provide a configuration of re-entry turbine that improves the efficiency by making use of pressure compounding, where the pressure drops in each of the stationary passages (entries) generate the appropriate flow conditions (velocity and angle) at entry to both inflow and outflow passes through the rotor blades whereby the impulse condition occurs and all blades in the flowpath experience near optimal incidence at the design point conditions. A form of pressure compounding is disclosed in DE311345, published in 1911.

[0006] However, despite this early disclosure, pressure compounding has not been widely employed in the design of re-entry turbines, one problem being that the pressure drop between re-entry stages can give rise to leakages around the circumference of the turbine disc. Conventionally, the solution to this problem is to eliminate the pressure drops by the use of velocity compounding.

Summary of the Invention

[0007] According to the invention, there is provided a radial re-entry turbine comprising a planar turbine wheel having a ring consisting of a plurality of turbine blades provided on at least one face of the wheel, the turbine wheel being mounted for rotation in a turbine housing having fluid flow passages therein providing at least two fluid entries into the or each ring of turbine blades, with the entries alternating between radial inflow and radial outflow or radial outflow and radial inflow, wherein the geometry of the fluid flow passages and the turbine blades is such that there is a pressure drop in each of the passages between successive arcs of blades but substantially zero pressure drop across each arc of blades in the ring of blades, characterised by at least one leakage blocking element configured to restrict circumferential leakage flow of the working fluid around the turbine wheel from a higher pressure turbine arc to a lower pressure turbine arc.

[0008] The turbine is thus configured to operate as a series of impulse turbine stages on a single wheel.

[0009] Other features of the invention are set out in the claims.

[0010] Typical applications of the turbine of the invention with compressible fluids are (a) in Rankine cycle machines for waste heat from smaller scale applications such as internal combustion engines, (b) smaller steam applications, for example with waste steam, and (c) with industrial gases.

[0011] Rankine cycle machines could conveniently use organic refriger- ant/air-conditioner gases such as R134A, or natural refrigerants such as carbon dioxide, enabling energy recovery at relatively low temperatures. The turbine of the invention could also be used for incompressible fluids, e.g. water, where power extraction is required from a high head supply.

[0012] The turbine of the invention may be used to drive an electricity generator, which may be coupled to the turbine shaft or may be constructed with a common shaft. However, the turbine could be used to provide mechanical power for use in driving other machines, for example powering industrial processes. [0013] Another aspect of the invention provides a radial re-entry turbine having at least two entries into a single turbine wheel, with entries alternating between radial inflow and radial outflow or radial outflow and radial inflow, characterised in that the geometry of the entries and the turbine blades is such that there is a pressure drop between each arc of blades but substantially zero pressure drop across each arc of blades in the turbine wheel, wherein the turbine blades have different leading-edge angles in the radial inflow and radial outflow directions such that there is near optimal incidence in both flow directions consistent with zero pressure drop in both flow directions.

Brief Description of the Drawings

[0014] In the drawings, which illustrate exemplary embodiments of the invention:

Figure 1 is a trimetric view of the exterior of the turbine assembly;

Figure 2 is a sectional view of the turbine to highlight different possible embodiments of the design;

Figure 3 is the sectional view A-A from figure 2;

Figure 4 is a detailed view B in Figure 2 of an embodiment in which the turbine inner stator plate and inner seal plate form a single component separate to the turbine outer stator plate;

Figure 5 is a detailed view B corresponding to that of Figure 4 of an embodiment of the invention where the inner and outer turbine stators are a single component, the turbine inner seal in this instance being a separate component;

Figure 6 is the detailed view B corresponding to that of Figure 4 illustrating the first instance of the invention with a single sided turbine wheel (bladed disc or blisk);

Figure 7 is the detailed view B corresponding to that of Figure 4 illustrating an embodiment of the invention with a double-sided turbine wheel (bladed disc or blisk);

Figure 8 is a detailed view of an embodiment of the invention with circumferential leakage blockers at key positions around the axial seals; Figure 9 is a section on C-C in Figure 8, showing the leakage blocker in the running position;

Figure 10 is a view corresponding to that of Figure 9, but with the leakage blocker shown in the retracted position to allow assembly;

Figure 11 is a view corresponding to Figure 3 illustrating the possible locations of the leakage blockers shown in Figures 8-10;

Figure 12 is an enlarged view of a portion of the turbine wheel of Figure 11 , illustrating the preferred blade configuration; and

Figures 13 and 14 show, respectively, examples of the velocity components in such a scenario for the inflow and outflow directions.

Detailed Description of the Illustrated Embodiment

[0015] Referring to Figure 1 , the turbine has a bearing section 1 and, at one end thereof, a turbine housing 2 provided on the end face with a fluid input 3 and a fluid exhaust 4. As may be seen from the cross-sectional view of Figure 2, the bearing section 1 has a bearing housing 5 in which a rotor shaft 6 is supported on two main bearings 7 and 8. The shaft 6 is terminated with a coupling 9 by which it can drive an electricity generator or mechanical device (not shown). The bearing housing 5 is also provided with a lubricating oil feed 10 and oil drain 11 .

[0016] The shaft 6 includes, at the end thereof remote from the coupling 9, a waisted extension 12 to receive the turbine disc 13, which is secured by a nut 14 engaging a threaded portion 15 of the shaft. The turbine housing 2 is secured to one end of the bearing housing 5 by screws 16 and encloses inner and outer turbine stator plates 17 and 18, with an outer seal plate 19 surrounding the turbine wheel.

[0017] Referring to Figure 3, which shows an end sectional view of the turbine disc 13 and stator plates 17 and 18, the fluid input 3 admits pressurised fluid to an entry cavity 20 of the outer stator plate 18. The entry cavity 20 directs the fluid flow through a first arc 21 of the ring of turbine blades 22 arranged around the turbine disc 13 to enter a passage 23 through the inner stator plate 17. The passage 23 directs the flow outwardly through a second arc 24 of the turbine blades 22 and into a second cavity or passage 25 in the outer stator plate 18, which is configured to direct the flow inwardly through a third arc 26 of the turbine blades 22 and into a cavity 27 in the inner stator plate 17, which cavity is in communication with the fluid exhaust 4. The cavities/passages 20, 23 and 25 are configured so that the fluid pressure drops within each passage but without significant drop through each arc of the turbine blades. The pressure drop in each of passages 20, 23 and 25 is such as to cause the fluid velocity presented at inflow to each arc to be appropriate to generate the no pressure drop condition in the turbine blades. The arc length of each admission is set whereby the resultant flow angle is such that it matches the leading-edge angle of the blades (minimum incidence condition).

[0018] Cascade vanes 28 are provided in the passage 23 in the inner stator plate 17 to guide the flow with minimum pressure loss. Similarly, a further set of cascade vanes 29 is provided in the passage 25 in the outer stator plate 18.

[0019] Figure 4 shows an enlarged sectional view of detail B of Figure 2. The inner stator plate 17 is separate from the outer stator plate 18 and incorporates the inner seal to the turbine disc. The enlarged view shows the pressure balance holes 31 provided between the opposed faces of the turbine disc 13 to minimise axial thrust. Figure 5, shows a corresponding view of an alternative embodiment, in which the inner and outer stator plates 17 and 18 are formed as a single component, with an inner turbine seal plate 30 being provided separately from the stator plate.

[0020] Figure 6 is an enlarged sectional view of the detail B of Figure 2 illustrating the sealing arrangements for the single-sided turbine disc employed in that embodiment. Two knife edge seals 32 and 33 are provided around the outer face of the turbine disc 13 and two further knife edge seals 34 and 35 surround the periphery of the inner seal plate 30. These can be integral with the disc/seal plate, although for large sizes they could be separate seal strips secured using caulking wire, for example. The knife edge seals serve to minimise leakage of the working fluid in the axial direction, i.e. from one face of the wheel to the other. It will be appreciated that, while knife edge seals are particularly suitable for use in this type of turbine, other seals, such as brush seals, could be employed alternatively or additionally.

[0021] Referring to Figure 7, showing an enlarged sectional view corresponding to portion B of Figure 2, but showing an alternative embodiment in which the turbine disc is double-sided, the fluid input being divided between the two faces of the disc in parallel flow configuration, or the outflow from one side of the disc becomes the inflow to the other side of the disc in series flow configuration, in which case additional seals may be introduced to the outer disc face.

[0022] Figure 8 shows an alternative arrangement to that shown in Figure 6. In order to reduce circumferential leakage in the passages bounded by the knife edge seals, the turbine disc and the seal plates, leakage blocker seal strips 36 and 37 can be provided to co-operate with the knife edge seals 32 and 33 respectively. The internal leakage blocker 36 is retractable to facilitate assembly of the turbine. Figures 9 and 10 illustrate the leakage blocker 36 in running and retracted positions respectively. Figure 11 shows possible locations for the leakage blockers at circumferential positions 40, 41 and 42 on the turbine wheel.

[0023] It will be appreciated that, while the embodiments described with reference to the drawings have flows start with an inflow, it would be possible to introduce the fluid within the inner stator plate so that the first stage is radial outflow, followed by radial inflow and then radial outflow again to exhaust through the outer stator plate. Also, the number of stages or passages of the fluid through the turbine ring is not limited to three, but there will necessarily be at least two.

[0024] While the invention has been illustrated by embodiments in which the flow is exhausted radially inwards from the turbine ring, it will be understood that the flow could equally be arranged to exhaust radially outwardly of the turbine ring and indeed radially outwardly of the turbine housing as well.

[0025] The fluid flowing through the turbine may be steam, but other gaseous fluids may be employed, such as organic refrigerant/air-conditioner gases, to enable energy recovery at relatively low temperature, as well as incompressible fluids such as water.

[0026] The turbine of the invention is suitably coupled to an electricity generator to enable the generation of electricity from waste heat sources, for example, but may be used for driving other power absorption/transfer devices, such as pumps and compressors. A gearbox may be used to adapt the speed of the turbine to the speed of the connected driven device.

[0027] Referring now to Figure 12, each turbine blade 22 comprises two surfaces 22a and 22b that come together at an edge at the inner and outer radius of the bladed region. The edge may be radiused or cut-off to avoid a sharp point. The camber line (dashed) is the line midway between the two surfaces. The angle a, [3, respectively, the camber line makes to a radial line at the outer or inner radius is defined as the leading edge angle for radial inflow or radial outflow respectively. The blade leading edge angles are defined such that the angle the incoming flow makes to the leading edge, the incidence, is optimal, i.e. generating minimum pressure loss. Typically, the optimal incidence angle is close to zero degrees. The relationship between the inflow and outflow leading edge angles is such that there is no significant difference in pressure at the inner and outer radius whether the flow is inflow or outflow in nature when the leading edge flow angle is at optimal incidence in both cases. The precise relationship is established, based on geometry and aerodynamic conditions according to a closed set of fluids equations including: conservation of rothalpy (rotational energy), conservation of mass and the isobaric enthalpy change across the turbine blades consequent to the aerodynamic energy loss. Referring to Figures 13 and 14, W is the relative flow velocity and VR is the radial velocity component. Wi and W2 are related by conservation of rothalpy, whereas VR1 and VR2 are related by mass conservation. Each radial velocity is equal to the respective relative flow velocity multiplied by the cosine of the flow angle.