CA2199781A1 - Hydraulic engines with at least two counterrotating runners - Google Patents
Hydraulic engines with at least two counterrotating runnersInfo
- Publication number
- CA2199781A1 CA2199781A1 CA002199781A CA2199781A CA2199781A1 CA 2199781 A1 CA2199781 A1 CA 2199781A1 CA 002199781 A CA002199781 A CA 002199781A CA 2199781 A CA2199781 A CA 2199781A CA 2199781 A1 CA2199781 A1 CA 2199781A1
- Authority
- CA
- Canada
- Prior art keywords
- runner
- exit
- turbine
- hydraulic engine
- shaft
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Abandoned
Links
Classifications
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04D—NON-POSITIVE-DISPLACEMENT PUMPS
- F04D3/00—Axial-flow pumps
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
- F01D—NON-POSITIVE DISPLACEMENT MACHINES OR ENGINES, e.g. STEAM TURBINES
- F01D1/00—Non-positive-displacement machines or engines, e.g. steam turbines
- F01D1/24—Non-positive-displacement machines or engines, e.g. steam turbines characterised by counter-rotating rotors subjected to same working fluid stream without intermediate stator blades or the like
- F01D1/26—Non-positive-displacement machines or engines, e.g. steam turbines characterised by counter-rotating rotors subjected to same working fluid stream without intermediate stator blades or the like traversed by the working-fluid substantially axially
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F03—MACHINES OR ENGINES FOR LIQUIDS; WIND, SPRING, OR WEIGHT MOTORS; PRODUCING MECHANICAL POWER OR A REACTIVE PROPULSIVE THRUST, NOT OTHERWISE PROVIDED FOR
- F03B—MACHINES OR ENGINES FOR LIQUIDS
- F03B13/00—Adaptations of machines or engines for special use; Combinations of machines or engines with driving or driven apparatus; Power stations or aggregates
- F03B13/08—Machine or engine aggregates in dams or the like; Conduits therefor, e.g. diffusors
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F03—MACHINES OR ENGINES FOR LIQUIDS; WIND, SPRING, OR WEIGHT MOTORS; PRODUCING MECHANICAL POWER OR A REACTIVE PROPULSIVE THRUST, NOT OTHERWISE PROVIDED FOR
- F03B—MACHINES OR ENGINES FOR LIQUIDS
- F03B13/00—Adaptations of machines or engines for special use; Combinations of machines or engines with driving or driven apparatus; Power stations or aggregates
- F03B13/10—Submerged units incorporating electric generators or motors
-
- Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
- Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
- Y02E—REDUCTION OF GREENHOUSE GAS [GHG] EMISSIONS, RELATED TO ENERGY GENERATION, TRANSMISSION OR DISTRIBUTION
- Y02E10/00—Energy generation through renewable energy sources
- Y02E10/20—Hydro energy
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- Engineering & Computer Science (AREA)
- Mechanical Engineering (AREA)
- General Engineering & Computer Science (AREA)
- Chemical & Material Sciences (AREA)
- Combustion & Propulsion (AREA)
- Hydraulic Turbines (AREA)
Abstract
A hydraulic turbine or pump without entry or exit guide vanes supported by a cantilevered stationary tube having a pair of concentric counterrotating first and second shafts with an exit section inclined at an angle relative to the axis of the shafts. A first runner is mounted to the first shaft, including at least a pair of radially extending first blades. A second runner is mounted to the second shaft and adapted for counterrotation relative to the first runner. The second runner has at least a pair of radially extending second blades. The pitch of the first and second runner blades is equal or different but inversed, where the runners have opposite, either equal or different, rotating speeds. A Y shaped tube housing encloses the turbine or pump to allow a withdrawal of the complete turbine or pump for inspection or maintenance.
Description
' 1 - 21 ~q7~1 HYDRAULIC ENGINES WITH AT LEAST
TnO COUNTERROTATING RUNNERS
The invention relates to a hydraulic engine with at least two counterrotation runners without en-trance and exit guide vanes. The hydraulic engine can be either a turbine or a pump.
Hydraulic turbines driving an alternator must be operated at optimal conditions, therefore at high rotational speed n (r.p.m.) to obtain low cost machines 0 and avoid highly geared generators. The number of sets for the same capacity should be small. Commercial alternators have a a standard speed n r.p.m. of either 1800 ~1500 or other) r.p.m. or 3600 (3000 or other) r.p.m. Increasingly, mass produced generators of 3600 (3000 or other) r.p.m. are used, since they are, for the same capacity, less costly and also less heavy than the 1800 (1500 or other) r.p.m. units. These 1800 (1500 or other) r.p.m. and particularly the 3600 (3000 or other) r.p.m. alternators have a high power/mass ratio which is in the same range as the power/mass ratio of the turbine.
Normally in hydraulic turbines of a certain power output, variations in the load are compensated by controlling the water flow through the runner. This is done by changing the position of the guide vanes and for axial flow turbines also the position of the runner blades. This classical solution, typical for Kaplan -and Bulb-Turbines, yields a high and flat peripheral or hydraulic efficiency ~h over the whole load range.
30 Small hydraulic turbines, equipped with adjustable vanes are very expensive and the cost of maintenance is not negligible. The cost of axial flow turbines can be substantially reduced by removing partially or even totally the adjusting vane mechanism.
TnO COUNTERROTATING RUNNERS
The invention relates to a hydraulic engine with at least two counterrotation runners without en-trance and exit guide vanes. The hydraulic engine can be either a turbine or a pump.
Hydraulic turbines driving an alternator must be operated at optimal conditions, therefore at high rotational speed n (r.p.m.) to obtain low cost machines 0 and avoid highly geared generators. The number of sets for the same capacity should be small. Commercial alternators have a a standard speed n r.p.m. of either 1800 ~1500 or other) r.p.m. or 3600 (3000 or other) r.p.m. Increasingly, mass produced generators of 3600 (3000 or other) r.p.m. are used, since they are, for the same capacity, less costly and also less heavy than the 1800 (1500 or other) r.p.m. units. These 1800 (1500 or other) r.p.m. and particularly the 3600 (3000 or other) r.p.m. alternators have a high power/mass ratio which is in the same range as the power/mass ratio of the turbine.
Normally in hydraulic turbines of a certain power output, variations in the load are compensated by controlling the water flow through the runner. This is done by changing the position of the guide vanes and for axial flow turbines also the position of the runner blades. This classical solution, typical for Kaplan -and Bulb-Turbines, yields a high and flat peripheral or hydraulic efficiency ~h over the whole load range.
30 Small hydraulic turbines, equipped with adjustable vanes are very expensive and the cost of maintenance is not negligible. The cost of axial flow turbines can be substantially reduced by removing partially or even totally the adjusting vane mechanism.
- 2 - 21 997~1 A layout with adjustable runner blades and fixed guide vanes or adjustable guide vanes and fixed runner blades, the latter called Propeller - turbine, is possible. The cost advantage is sharply off-set by the very peaked efficiency - power curve owing to the sensitivity of axial flow blades to incidence and is even more pronounced for Propeller turbines. This holds particularly for small heads.
Radially arranged guide vanes in axial flow turbines have straight exit edges, parallel to the turbine shaft. Tight closure of the turbine is possi-ble. The guide vanes of axial flow turbines in a straight water duct require a twist, needed for the generation of a constant moment of momentum vs. radius before the runner. Such guide vanes must be carefully matched to the runner blades and are therefore costly to fabricate. The tight closure of the turbine is impossible. To obtain satisfying flow control opera-tion, a large fly-wheel mass is therefore required to 20 stabilize the speed control and to limit the overspeed after a sudden load rejection or when a safety parame-ter, constantly monitored, initiates a shut down proce-dure. In case of water flow control, maximum closing and opening speed causes water hammer effects and tends to destabilize the control sequence. The closing speed results from the admissible pressure surge due to water hammer in the piping and possible water separation in the draft tube. These conditions can make a promising development too expensive.
If a low power output hydro-electric set can be connected to the public AC grid a frequency control system can be dispensed of and only a guide vane dis-placement mechanism may be required, but an emergency shut down mechanism is necessary. If a hydraulic tur-bine supplies a grid by itself, the frequency will be f for any power generated between PmaX and Pmin and is fmax for Pmin and fmin for Pmax, the relative frequency ~1 ~ j7d 1 deviation corresponding to maximum and minimum power consumption is then represented by ~P = (f G max ~ f G min ) / f G nominal called the speed droop which, under favorable condi-tions, is approximately 5% or 7% or higher. The sub-script G indicates the grid fG rated=(fG max+fG min)/2 This expression gives only the steady state speed settled deviation and not the temporary speed variation of a control sequence. Dynamic stability of a control sequence depend on the moment of inertia of the set, the speed self - control of different kinds of load, the load degree and the length and shape of the penstock. As a general rule, a large power hydro turbo set and its alternator has to withstand the runaway speed n ra of the turbine. The value of n ra depends on the design of the turbine and its components and on the head H. The runaway speed of axial flow turbines with small guide vane angles a and small runner blade angles ~ between relative velocity and peripheral blade 20 speed can be in the range of (1,4 to 3,3 ) n rated.
Constructive means, reserved for larger power output axial flow turbines can reduce a high runaway speed by a suitable runner blade profile to introduce an opening tendency of adjustable runner vanes, thus increasing ~ once the unit is disconnected from the grid or, when using step-up planetary gears, by discon-necting the outside gear housing from the ground. Other sophisticated systems also exist If a hydraulic turbine supplies a grid by 30 itself and a constant frequency f G rated of 60 (50 or other) Hz is required the unit should be operated by load-control.
Contrary to custom-built alternators to with-stand a runaway speed, inexpensive, mass produced alternators of 1800 (1500 or other) r.p.m. or 3600 (3000 or other) r.p.m. can withstand only an overspeed 7 ,, i of 1,3 n rated except units for very small output, where the overspeed can be very hlgh. A very effective fail-proof brake down system must be provided to shut down the set.
The invention has the object to provide a hydraulic engine having two counterrotating runners without entrance and exit guide vanes which has a small constructional length, can operate under a head range not possible for one stage unit and can provoke a lower o runaway speed. The head H is divided equally or unevenly between the counterrotating runners. Further-more the hydraulic engine should be manufactured at low costs with each runner having at least two blades each twisted in the circumference of the corresponding run-ner, whereby the blades of the entrance runner are twisted in the opposite direction relative to the blades of the exit runner.
The engine with its counterrotating runners is of simple construction, can harness a head H at a 20 positive suction head h5 for which a one stage unit cannot be designed. The layout of a turbine with coun-terrotating runners is greatly simplified since no guide vane runner blade match is necessary and the constructional length can be small. Due to the simple construction, the engine is practically maintenance free or of low maintenance. In the case of a turbine it can drive a low cost high speed alternator or several alternators of the same rotational speed to furnish the rated or requested constant frequency f. To obtain a 30 low lift coefficient ~A of the runner blades, the engine runners must rotate at high speed n r.p.m.
These and further objects will be more read-ily appreciated when considering the following disclo-sure and appended drawings wherein:
Fig. 1 is a schematic view which shows an axial flow one stage hydraulic turbine accordlng to the state of art;
~ 1 9~& 1 Fig. 2 is an axial cross-section showing a hydraulic turbine in accordance with the present invention in longitudinal cross-section;
Fig. 2a is a radial cross-section showing a blade of a runner;
Fig. 2b is a schematic view of a blade cascade of runner R1 and runner R2 unrolled in a plane;
Figs. 3a and 3b are diagrams showing the velocity triangles of the turbine of Fig. 2;
oFig. 4 is a view of a camfered thin circular arc blade of the present invention;
Fig. 5 is a side elevation partly in cross section of an alternator in accordance with the present invention;
Fig. 6 is a vertical cross section taken along line A-A of Fig. 5;
Fig. 6a is an enlarged exploded view of a detail shown in Fig. 6;
Fig. 7 is a vertical cross section taken 20 along line B-B of Fig. 5;
Fig. 8 is a schematic view of an alternator stator and rotor coupling in accordance with the pre-sent invention;
Fig. 9 is an axial cross-section showing a hydraulic pump; and Figs. lOa and lOb are diagrams showing the velocity triangles of the pump of the present inven-tion.
A known low cost turbine 1 as shown in Fig. 1 has a housing 2 composed of a long or regular elbow and a flared exit section and has no entrance nor exit guide vanes and only the runner blades 4 mounted on the runner hub 3 assure the flow deflection. The water flow before the entrance edge EB1 and behind the exit edge EB2 of the blades is irrotational and submitted to the potential law, characterized by cur = constant ~1 ~ 7 7~ i with r the radial distance of a water particle from the turbine shaft and CU its whirl component. With reference to Fig. 1, cul = O and CU2 is the whirl component behind the blades 4 at various cylindrical flow sections is reduced to cu5 in the draft tube, composed of the flared housing 2 and the coaxially arranged cone 9. The symbols c and cm and cmS stand for the meridional water velocities before the turbine, through the runner and in the draft tube exit section.
o The turbine 1 is mounted in a cantilever manner with shaft 5 journaled in a bearing housing 6.
The shaft 5 drives an alternator 8 by means of gears 7 or pulleys and straps.
The rotational flow in the energy recuperat-ing flow duct, called the draft tube, allows a very high diffusor angle. The radial spread of the cone 9 must sharply increase to assure an effective reduction of CU2 and to reduce friction or energy losses in the draft tube. The energy losses between the surfaces of 20 the inner cone 9 and the bell shaped turbine housing can become high, thus lowering the efficiency of a long flow guiding element. For low heads H, the energy recuperation in the draft tube from cu2 to cu5 can give satisfying results. If a certain power necessitates a high flow volume Q, the turbine 1 will be of rather large dimensions. With increase of head H, the high whirl component CU2 at the blade exit edge EB2 can only be reduced to a small cu5 by a cone 9 of severe expan-sion and axis-symmetric velocity distribution will no longer exist. When the turbine 1 operates away from the best efficiency point, at for example varying head H, the presence of a vortex core, called cork screw in addition to a whirl results in non axis-symmetric flow.
This will result in a stall and boundary layer separa-tion in the draft tube reducing furthermore its effi-ciency.
_ 7 _ 21 ~7~, Since a stall can lead to vibrations, such a cone equipped draft tube must be of a very sturdy construction. Eor this reason, cones made of concrete cannot be used, since the irregular flow will demolish surface sections and the cone is ultimately washed away.
The turbine should have by preference a ver-tical shaft. The overall efficiency 1l of such a tur-bine 1 will be rather low and depends on the reduction o of cu2 to cus An energy loss resulting from a whirl component cus at the blade 4 exit edge EB2 must be carefully weighted against the whirl component at the draft tube exit and must be added to the friction losses of the cone-equipped bell shaped draft tube.
This will determine if such a set having a turbine without guide vanes is economically viable.
The embodiment of the invention shown in Fig. 2 relates to a hydraulic turbine lO which is supported by a stationary cantilevered hollow tube 12 20 fixed to the housing plug 14 and extends in turbine casing 11. The casing ll is in form of a Y with equal cylindrical inner diameters in smooth communication.
The casing 11 includes a coaxial first tube leg llA and a second casing tube leg llB at an angle to the axis of the hollow tube 12 set in casing tube leg llA. The inlet and outlet sections are inclined at an angle towards the stationary hollow tube 12. The turbine 10 has no entrance nor exit guide vanes and has two counterrotating runners R1 and R2. Each of the rum~ers 30 R1 and R2 has at least two blades P1 and P2 of equal outside and hub diameter and a reversed pitch which is of equal or different value. The entrance runner R1 is arranged on a rotating solid shaft Sl coaxially mounted in the counterrotating hollow shaft S2 which supports the exit runner R2. The runner R2 includes a bearing 20 mounted to the shaft S1 and a bearing 21 mounted to the stationary tube 12. The shaft S2 is journaled in bearing 15 set in the housing 14. The shaft S1 is journaled in the hollow shaft S2 and supported by bearing 20 and bearing 17.
The mechanical seal assembly 19 (or other suitable seal) excludes water entering the space between the hollow rotating shaft S2 and the solid shaft S1. The mechanical seal assembly 13 (or other suitable seal) separates the turbine 10 from the ambient atmosphere. The entrance hub 22 screws onto the shaft S1 with counter threads to the rotation shaft S1 so that the hub will always be firmly tightened against spacer 25. The exit runner hub 24 is keyed to the hollow shaft S2 and tightened by an outside threaded insert ring 26 with an inside split conical bushing.
Pulley BP'R2 is threaded onto shaft S2. The shaft S2 is journaled in bearing 15 set in the housing plug 14. The shaft S1 is supported at its exit end by bearing 17 carried by pulley BP'R2. The pulley BP'R1 is directly connected to the shaft S1. Housing plug 14 is 20 firmly retained in casing tube leg llA. A machine bolt 23 is shown fastening plug 14 to the casing 11. The interior diameter of tube leg llA is at least greater than the diameter of the runners R1 and R2 so that the housing plug 14 and the turbine 10 can be removed through the opening formed at the end of casing tube leg llA.
By this arrangement, the complete turbine can be withdrawn for inspection of repair purpose or for storing during interruption of activities.
The blades Pl and P2 are made for utmost simple fabrication from flat plate material. Once traced and cut, the outside section E5 of blade P1 is twisted afterwards relative to the hub section E6 in such a way that E5 and E6 remain flat. The flat blade foot FE6 is inserted in the straight slots 16 of hub 22 of Fig. 2 and then screwed, welded or glued and after-wards machined to the outside diameter tolerances.
~7 ~7 ' Blade P2 with blade foot FE7 (not shown) is fabricated and installed in a similar manner. Fig. 2a is the view of a blade P1 in the direction of the axis a-a and Fig. 2b is the view normal to the turbine axis a-a of the blade cascade P1 and P2 unrolled in a plane.
Twisting the blades P1 and P2 must exclude a spring effect during manufacturing or under load. The blades of non rusting material should be of a short radial length to keep the twist between outside and inside sections small and thus the bending arm towards the hub is also short. For shockless entry the entrance edges of the rotating flat blades P1 and P2 are elliptically rounded off. Blunt trailing edges are not recommended and the blades should be sharpened at the exit edge E3 and E4. If special conditions require, arc curved blades of equal thickness (Fig. 4) or profiled ones must be used, but they are generally more complicated to fabricate. The entrance edges E1 and E2 of the runner blades P1 and P2 are either in the radial 20 direction or skewed at a small angle with the radius in which case the exit edges E3 and E4 will also be skewed. The rotor R1 has a rotational speed n Rl and a circumferential velocity Urunner R1, the rotor R2 has a rotational speed n R2 and the circumferential velocity Urunner R2 with n R1 and n R2 being equal or different.
The water velocity before blades P1 and behind blades P2 is c and the relative water velocities in the water channels are wp1 and wp2. The acute angle of the slots 16 and 18, (~0O-~) in Fig. 3 and the twist depend on 30 the working data H, Q and n.
In conventional one-stage axial flow turbines the guide vanes must be welded along their contour to the bearing housing and to the turbine casing and a welding gauge is required. The radial bearing must then be carefully centered and this requires much manual work.
~1 ~ 7 I
In contrast to this, the fabrication of the turbine 10 with counterrotating runners R1 and R2 is simple and no guide vanes are required.
Reference is now made to Figs. 3a and 3b. A
fluid particle on its way through the turbine runners R1 and R2 remains on the surface of a cylinder, an assumption which holds true for hydraulic efficiencies ~h close to the maximum value. A blade P1 and P2 is unrolled in a plane, as Fig. 2b shows. In the absence o of whirl CU the absolute velocity cO is then cm. This Cm is resolved due to the blade peripheral speed UrunnerRl into the relative velocity w1 at the entrance edge E1 of blade P1 and the relative velocity w2 at the runner exit edge E3 produces together with Urunner the absolute velocity C3 between the runner R1 and R2.
The relative velocity wooR1 exists for a relative flow before and behind the blade P1 and allows together with the deflection ~CU=C3U - cOU
20 the calculation of the lift coefficient ~A.
The absolute velocity C3 is resolved due to the runner R2 with UrunnerR2 into w4 at the entrance edge E4 and is then reconverted due to Ws and UrunnerR2 into C6 = cm. The velocities UrunnerR1 and UrunnerR2 need not be equal. The relative flow in a channel of runner R1 and runner R2 is wp1 and wp2.
If blade P1 is arranged in the direction of WooR1 its ~A = ~ and no power can be generated. A
physical angle of attack ~R1 between the flat blade and 30 the relative flow wooRl produces w1 and w2, ~A takes up positive values, producing the forces A, Au and AaX and the turbine generates power. For curved flat plate skeletons (Fig. 4) or profiled blades the basic relations holds also but are more complicated.
To avoid cavitation the absolute velocity Cm through the turbine will be kept small and the turbine composed of the runner R1 and the runner R2 can be placed on solid ground above the highest tail water level. The danger of an accident or an electric shock is greatly reduced but strict security measures must be respected carefully and scrupulously.
In the embodiment of Fig. 4 a blade P1 or P2 with a circular arc profile is shown in a schematic view with a radius ra. The length L (Fig. 4) of the runner blades P1, P2 is of importance. The wakes of long runner blades L dissipate more slowly than those o of shorter ones, whereas a large 1 leads to a small lift coefficient ~A. If a turbine runaway occurs, runaway speed will be rather low due to the highly disturbed flow field behind the blades P1 in which the exit counterrotating runner blades P2 wade. Since the two runners R1 and R2 are mechanically coupled to the alternator rotor the runaway speed of such turbines is inferior to that of a one stage axial flow turbine and for certain generators no mechanical braking is required and the station is shut down by stop logs or a 20 simple butterfly valve arranged at the penstock entry.
In the draft tube D of an axial flow tur-bine 10 (Fig. 2) with two counterrotating runners R1 and R2 without entrance and exit guide vanes, the kinetic energy lost is that of the water flow at the final turbine discharge cross section. This energy depends on cm2/2 and can necessitate the installation of a draft tube which will be of simple geometric con-struction, either flared or consists of a pipe with a sudden enlargement of the turbine exit diameter.
The operation of a plant consisting of a turbine 10 with two counterrotating runners R1 and R2 and its generator at variations of the head H will now be described. For load control the frequency f of the autonomous AC grid remains constant and the rotational speed of the runners R1 and R2 therefore keeps its constant value, irrespective of the operating value of H. For an actual head H the strict relation - 12 - 2i ~
llhgH = u ~CU
dictates for a blade speed u the value of ~CU and the components of the absolute velocities cO and C3 of the runners R1 and R2 (Fig. 3).
In the event the head H increases, the blade speed u is kept constant by the frequency control sys-tem. and the deflection ~CU must therefore lncrease.
Since the blade angle (Boo-~) (see Fig. 3) cannot be altered, an increase in ~CU can only be obtained by a o decrease of cm/ the meridian water velocity through the runner resulting in a reduction of the water volume flow Q. The unavoidable presence of shock losses will lower the hydraulic efficiency value of the turbine with two counterrotating runners. The power output of the turbine can therefore be inferior to the power of a similar turbine designed for this higher value H.
In the event the head H decreases, the blade speed u is kept constant by the frequency control sys-tem and the deflection ~CU must therefore decrease. A
20 decrease in ~CU can only be obtained by an increase of Cm the meridian water velocity through the runner ~Fig. 3) resulting in an increase of the water volume flow Q. It could be possible that ~CU becomes so small that the increase in cm and therefore of the volume flow Q through the rotor plate cascade suffers little deflection. A sharp decrease in efficiency will override the tendency to increase the volume flow and the turbine can produce less or no power at all.
When there are variations of the head H and 30 of frequency f with the turbine and its asynchronous generator connected to an existing AC grid having small variations of the frequency f, the grid linked fre-quency f of the asynchronous generator is within the speed droop ~p and will be for a large grid 1% to 2Q6, for a smaller one 5% or 7% or higher. The rotative velocities of the set are insignificantly modified and - 13 - 2i ~7 7~i the afore conclusions remain therefore valid. The power output will be slightly altered.
The blades P1, P2 of the turbine are posi-tioned for the rated values of the head H and volume flow Q. For a downward variation of the head H, the stilling basin must have some spill or a skip arrange-ment limiting its lowest head water level H minimum and allowing only an upwards varying head water level and the turbine layout must be made for the lowest value o of H.
In the event the rated head H of the turbine is very high, the lift coefficient ~A of blades P1 and blades P2 increases and the angle of incidellce ~ will be larger. A satisfying solution is that the runner R1 takes up a smaller head HR1 than the runner R2 which takes up a higher head HR2. This is achieved when the circumferential speed u of runner R1 is lower than the circumferential speed of runner R2. Since 71hgH = U~Cu 20 and UrunnerRl < UrUnnerR2 therefore HrunnerR1 < HrUnner R2 with HrunnerR1 + Hrunner R2 = H. Due to the different transmission ratios srunnerRl and Srunner R2 the power of runner R1 and runner R2 are cumulated in the mechanical alternator drive. The exit edge El of runner blades P1 shed a small vortex sheet and the stronger vortex sheet of runner R2 dissipates in the draft tube D without shock losses.
The rotor of the alternator A (Fig. 5) is the carrier of the magnetic field and is coupled with the turbine 10. For a small hydraulic set supplying its own autonomous grid, synchronous alternators must be used.
The alternator A is preferably single phased. The maxi-mum overspeed of the alternator A is about 1.3 of a rated alternator. An effective emergency shut down device such as a mechanical brake or a by-pass outlet - 19 ~ 7~ i can be required. Asynchronous generators can support a high runaway speed.
If the turbine is separated from its grid, the highly disturbed flow field of the two counter-rotating runners R1 and R2 will effectively limit the overspeed.
The turbine 10 can drive a low cost high rotational speed generator, i.e. alternator A or several generators of the same rotational speed.
0 The runners R1 and R2 can be coupled to the alternator A by different mechanical elements, such as bevel gears, by a gear arrangement, by a spur gear and a timing belt pulley, or by two timing belts. It is also possible to have the alternator rotor RT and the alternator casing ST directly coupled to the counter-rotating shafts S1 and S2 of the runners R1 and R2 (Fig. 8).
The greatest flexibility for coupling the turbine 10 to the alternator A is provided by the use of two timing belts (Figs. 5 to 7). On the shaft SA
belt pulley BP1 is coupled by the double gear belt BTD
with a belt pulley BP'R1 on the shaft S1. The diameter and the distance of the idler belt pulley BP3 determine the number of teeth in the mesh of the belt pulley sP1.
The rotation of the runner shaft S2 is transmitted to the alternator A by the pulley PB'R2 and the pulley BP2.
The transmission relations are either HrUnnerRl = HrunnerR2 SRRl = SRR2 HrUnnerRl + HrUnnerR2 = H
30 or HrUnnerRl i~ HrUnnerR2 SRRl ~ SRR2 HrUnnerRl + HrUnnerR2 = H
The conventional speed governor necessary for volume flow control cannot be used for the inventive turbine since the runner plates P1, P2 are of fixed geometry. For units connected to an existing grid the speed control is omitted. If the station supplies a -1S- ~ly9~
grid by itself, the frequency is kept constant by load control. For this operation the link between the tur-bine-alternator set produced power PA, the industrial load Pindustrial and the frequency control load Pfrequ is expressed by PA Pindust+Pfrequ Any difference between the available PA and the continuous variation of the industrial load Pil~dUst results in a frequency deviation tendency of the alter-o nator A. The continuously monitored alternatorfrequency is compared with the very accurate quartz reference frequency and the Pfrequ load, composed of suitable induction free resistors, is switched at a high speed across the alternator terminals. The power balance is observed and the frequency f remains con-stant. These resistors can be used for secondary indus-trial uses, such as water heating, drying, etc. This system is always dynamically stable, irrespective of the self control of various kinds of load. The total 20 dissipating capacity of the load frequency control system must be greater than the nominal power available at the alternator terminals to avoid that the frequency f goes up. If the turbine is overloaded the frequency f will go down. For a judicious partition between primary and secondary industrial load the overall operational efficiency which is of no dominant concern can nevertheless be high. The electronic security system, protecting the turbine and the alternator A keeps the various components under continuous surveillance and 30 limiting values of various parameters provoking an emergency shut-down.
Such a very sensitive frequency control sys-tem together with a safety system is patented (Canadian patent 1,244,081 - November 1, 1988 - Patent holder:
Dr.-Ing. H. Netsch).
- 16 - ~ ?~
The described turbine may be installed pref-erably for users far away from the electric public grid but can also be connected to an existing AC grid.
In summary the turbine can be equipped with a draft tube which can harness very low and very high heads with a positive suction head capable to operate at head fluctuations with a positive suction head avoiding thus turbine submergence and uneconomical civil engineering work. The turbine can be installed o with horizontal, vertical or inclined shaft and is practically maintenance free or of low maintenance. The turbine operates under very high heads for which one stage axial flow turbines cannot be developed and divides the available head potential energy evenly or unevenly between the two runners Rl and R2, so that the runner R1 at the turbine entrance produces a small power. The exit whirls of runner Rl are then of very small magnitude whereas the following counterrotating runner R2 produces a higher power with stronger exit 20 whirls dissipating in the draft tube with the absence of shock losses resulting in high hydraulic efficiencies. The power is cumulated in the mechanically coupled alternator. The turbine has a low runaway speed since the flow field of both runners R1 and R2 becomes very irregular with increased rotational speed n. In addition, the counterrotating runner R2 wades in the irregular flow field of the entrance runner R1. For certain types of generators a braking system can be omitted and falling stop logs or an~0 inexpensive butterfly valves shuts down the water flow.
The turbine is coupled by gears or a timing belt system to one or more high speed inexpensive generators and can be coupled directly to the counter-rotating stator and rotor of the generator. High fre-quency accuracy of said turbine with counterrotating runners supplying its own grid is assured by load con-trol (Canadian patent l,24~,081 - November 1, 1988 -- 17 - ~ 1 Y ~
Dr. Ing. H. Netsch), producing stable operation inde-pendent of the nature of the load. The frequency regu-lator has a satisfying degree of overall efficiency for skillful repartition of the load. The turbine generator set can be connected directly to an existing AC grid, operating thus with the frequency of the grid, dispens-ing with the frequency regulator, requiring only an emergency shut down system. Since the delivered power remains constant the hydraulic efficiency remains high, o also for frequency variations of the existing AC grid, since all flow velocities in the turbine are equally altered.
The complete cantilevered turbine 10 with two counterrotating runners R1 and R2 can be withdrawn for inspection or repair purpose or for storing during interruption of activities.
When blocking pulley BP'R1 and causing the then stationary blades P1 to become guide vanes, only the exit runner R2 rotates. The one stage cantilevered 20 turbine operating in the Y bend casing 11 can com-pletely be withdrawn.
The hydraulic engine of the present invention may also be in the form of a hydraulic pump with two counterrotating rotors without entrance and exit guide vanes. Hydraulic centrifugal or axial flow pumps should be operated at high rotational speed n r.p.m. and are coupled to high speed synchronous, or to less expen-sive, asynchronous motors of 3600 r.p.m. (3000 or other). For a certain power these 3600 r.p.m. motors are less costly and less heavy than units of lower speed and consequently the driven pumps are of smaller dimension. Centrifugal pumps have a high delivery head H and a small volume flow Q whereas axial flow pumps have a low delivery head H and a large volume flow Q.
An axial flow pump having two or more counterrotating rotors, without entrance and exit guide vanes, can have, for the same volume flow Q, a higher delivery - 18 - ~I Y~7~, head H (which a one stage axial flow pump cannot gener-ate) and has a small constructional length and is not costly to fabricate (Fig. 9).
The hydraulic axial flow pump 50 shown in Fig. 9 with two counterrotating rotors has no entrance or exit guide vanes, and has a casing 51 in the form of a Y bent with equal cylindrical inner diameters with inlet and exit sections inclined at an angle towards the stationary hollow cantilevered support tube 52 o which is fixed to the housing 54. As described in relation to Fig. 2, the Y shaped casing 51 has a tube leg 51A that is coaxial with tube 52, and a tube leg 51B at an angle to the axis of tube 52. The interior diameter of tube leg 51A is greater than the diameter of rotors PP1 and PP2 so that the plug 54 and pump 50 can be easily removed from the casing 51 through the tube leg 51A. The driven rotors PP1 and PP2 have at least two blades RP1 and RP2 of equal outside and inside diameters. Inside the hub 74 of the rotor RP2 is 20 a bearing 70 mounted to the driven shaft SP1 and a bearing 71 mounted on the stationary tube 52. The mechanical seal assembly 59 (or other suitable seal) excludes water entering the space between shafts SP1 and SP2 and the mechanical seal assembly 53 (or other suitable seal) separates the pump from the ambient atmosphere. The hollow shaft SP2 is journaled in bearing 55 set in the housing plug 54. The shaft SP1 is supported by bearing 57. The entrance hub 72 is screwed on the driven shaft SP1 which is threaded in the 30 countersense of rotation so that the hub 72 will always be firmly locked. The exit rotor hub 74 is keyed to the hollow shaft SP2 and locked by an outside threaded ring 76 having a split conical insert. The driven exit rotor RP2 is arranged on a hollow shaft SP2 of motor MP2 surrounding coaxially the shaft SP1 of motor MP1. Not excluded are driven systems consisting of spur gears, bevel gears or by two timing belts, as shown in Fig. 5 - 19 21 q',/,j j to 7. The blade PP1 of entrance rotors RP1 has an entrance edge EP1 and an exit edge EP3 and the blade PP2 of the exit rotor PR2 has the entrance edge EP2 and the exit edge EP4.
As described in relation to the turbine 10 of Fig. 2, the pump 50 can be removed with the plug 54.
Plug 54 is firmly installed in tube leg 51A by means of machine bolt 73. The plug 54 can be removed with tube 52 and pump 50 through the opening at the end of tube o leg 51A.
Thus for inspection, maintenance, repair or for operation interruption, the whole cantilevered pump can be withdrawn from its housing 51.
The blades are made for utmost simple fabrication from flat plate material and after tracing and cutting, the outside section of blade PP1 is then twisted relative to its hub section in such a way that these sections remain plane. For the blades PP2 of rotors RP2 the process is similar. The blades are made 20 from non rusting material. The foot of a blade is inserted into slots 56 or 58 and welded, screwed or glued to the respective hub 72 or 74. The blades should be short with elliptically rounded off entrance edges having no blunt trailing edges.
In Fig. lOb the velocity triangles for a radius r of the rotating and counterrotating blades PP1 and PP2 of an axial flow pump are shown. ~o entrance whirl exist, therefore CU = ~ and the absolute water flow velocity is Cpm which is resolved due to urOtorRpl 30 into the relative velocity wp1 at the entrance edge EPl of blade PP1 and the relative exit velocity wp2 at its exit edge EP3 produces the absolute velocity Cp3 between the rotor RP1 and RP2. This absolute velocity cp3 is resolved due to UrotorRP2 into wp4 and ~p5 together with urOtorRp2 produces cp6 = cpm(Cu = ~) the absolute water velocity at the exit of the pump.
Fig. lOa shows the blades, unrolled in a place similar - 20 - ~' I Y ~1 / () j to Figs. 3a and 3b. The velocity ~CpU together with WooPR1 allow the calculation of the lift coefficient ~A
and the inclination angle ~PR1 necessary to transmit the forces A, Au and AaX to the blade PP1. For generating a higher pressure head H the blades PP1 and PP2 of equal thickness are cambered in the form of a circular arc (Fig. 4) or profiled. To avoid priming of the pump the blade PP2 must be partially submerged otherwise a priming device must be provided.
o If the delivery head H is excessively high the blades PP1 and PP2 will shed wakes and the flow condition of the rotor RP2 will be disturbed. The following solution will correct this problem. The rotor RP1 generates a smaller pressure head HrotorRl whereas the rotor RP2 generates a larger pressure head HrotorR2 -delivery head HrotorRl < delivery head HrotorR2 Wakes shed by the plate exit edge EP3 will be of small magnitude and stronger wakes shed by the exit 20 edge EP4 of rotor R2 will dissipate in the flow without shock losses. The head HrotorRl and HrotorR2 will add up requiring a different power of motor MP1 and MP2. A
diffusor of suitable shape converts kinetic energy at the rotor exit to pressure.
If the volume flow Q decreases from its rated-value Q and constant speed is maintained, the head H increases and the efficiency falls off and the driving motors will be overloaded, as is also the case for single stage axial flow pumps. A stilling basin and 30 the high reservoir should be provided. Throttling valves should not be installed so that no markedly increase of the pressure head H and overloading the motor will occur. This very unfavorable operational condition which characterizes, also, one stage axial flow pumps, can be attenuated for pumps with counterrotating blade cascades if the variable speed - 21 - ~ l Y i /~ I' motors MP1 and MP2 are installed and operated in such a way that the delivered volume flow Q and the delivery head remain close to the rated design values of A
and H.
Back flow from the high to the low reservoir in case of shutdown or power failure can be avoided by a siphon in the pressure duct with a suitable aeration device.
When blocking shaft SP1, and causing the o stationary blades RP1 to become guide vanes, only the exit rotor RP2 rotates. The one stage pump operating in the Y bend casing 51 can completely be withdrawn through leg 51A.
Radially arranged guide vanes in axial flow turbines have straight exit edges, parallel to the turbine shaft. Tight closure of the turbine is possi-ble. The guide vanes of axial flow turbines in a straight water duct require a twist, needed for the generation of a constant moment of momentum vs. radius before the runner. Such guide vanes must be carefully matched to the runner blades and are therefore costly to fabricate. The tight closure of the turbine is impossible. To obtain satisfying flow control opera-tion, a large fly-wheel mass is therefore required to 20 stabilize the speed control and to limit the overspeed after a sudden load rejection or when a safety parame-ter, constantly monitored, initiates a shut down proce-dure. In case of water flow control, maximum closing and opening speed causes water hammer effects and tends to destabilize the control sequence. The closing speed results from the admissible pressure surge due to water hammer in the piping and possible water separation in the draft tube. These conditions can make a promising development too expensive.
If a low power output hydro-electric set can be connected to the public AC grid a frequency control system can be dispensed of and only a guide vane dis-placement mechanism may be required, but an emergency shut down mechanism is necessary. If a hydraulic tur-bine supplies a grid by itself, the frequency will be f for any power generated between PmaX and Pmin and is fmax for Pmin and fmin for Pmax, the relative frequency ~1 ~ j7d 1 deviation corresponding to maximum and minimum power consumption is then represented by ~P = (f G max ~ f G min ) / f G nominal called the speed droop which, under favorable condi-tions, is approximately 5% or 7% or higher. The sub-script G indicates the grid fG rated=(fG max+fG min)/2 This expression gives only the steady state speed settled deviation and not the temporary speed variation of a control sequence. Dynamic stability of a control sequence depend on the moment of inertia of the set, the speed self - control of different kinds of load, the load degree and the length and shape of the penstock. As a general rule, a large power hydro turbo set and its alternator has to withstand the runaway speed n ra of the turbine. The value of n ra depends on the design of the turbine and its components and on the head H. The runaway speed of axial flow turbines with small guide vane angles a and small runner blade angles ~ between relative velocity and peripheral blade 20 speed can be in the range of (1,4 to 3,3 ) n rated.
Constructive means, reserved for larger power output axial flow turbines can reduce a high runaway speed by a suitable runner blade profile to introduce an opening tendency of adjustable runner vanes, thus increasing ~ once the unit is disconnected from the grid or, when using step-up planetary gears, by discon-necting the outside gear housing from the ground. Other sophisticated systems also exist If a hydraulic turbine supplies a grid by 30 itself and a constant frequency f G rated of 60 (50 or other) Hz is required the unit should be operated by load-control.
Contrary to custom-built alternators to with-stand a runaway speed, inexpensive, mass produced alternators of 1800 (1500 or other) r.p.m. or 3600 (3000 or other) r.p.m. can withstand only an overspeed 7 ,, i of 1,3 n rated except units for very small output, where the overspeed can be very hlgh. A very effective fail-proof brake down system must be provided to shut down the set.
The invention has the object to provide a hydraulic engine having two counterrotating runners without entrance and exit guide vanes which has a small constructional length, can operate under a head range not possible for one stage unit and can provoke a lower o runaway speed. The head H is divided equally or unevenly between the counterrotating runners. Further-more the hydraulic engine should be manufactured at low costs with each runner having at least two blades each twisted in the circumference of the corresponding run-ner, whereby the blades of the entrance runner are twisted in the opposite direction relative to the blades of the exit runner.
The engine with its counterrotating runners is of simple construction, can harness a head H at a 20 positive suction head h5 for which a one stage unit cannot be designed. The layout of a turbine with coun-terrotating runners is greatly simplified since no guide vane runner blade match is necessary and the constructional length can be small. Due to the simple construction, the engine is practically maintenance free or of low maintenance. In the case of a turbine it can drive a low cost high speed alternator or several alternators of the same rotational speed to furnish the rated or requested constant frequency f. To obtain a 30 low lift coefficient ~A of the runner blades, the engine runners must rotate at high speed n r.p.m.
These and further objects will be more read-ily appreciated when considering the following disclo-sure and appended drawings wherein:
Fig. 1 is a schematic view which shows an axial flow one stage hydraulic turbine accordlng to the state of art;
~ 1 9~& 1 Fig. 2 is an axial cross-section showing a hydraulic turbine in accordance with the present invention in longitudinal cross-section;
Fig. 2a is a radial cross-section showing a blade of a runner;
Fig. 2b is a schematic view of a blade cascade of runner R1 and runner R2 unrolled in a plane;
Figs. 3a and 3b are diagrams showing the velocity triangles of the turbine of Fig. 2;
oFig. 4 is a view of a camfered thin circular arc blade of the present invention;
Fig. 5 is a side elevation partly in cross section of an alternator in accordance with the present invention;
Fig. 6 is a vertical cross section taken along line A-A of Fig. 5;
Fig. 6a is an enlarged exploded view of a detail shown in Fig. 6;
Fig. 7 is a vertical cross section taken 20 along line B-B of Fig. 5;
Fig. 8 is a schematic view of an alternator stator and rotor coupling in accordance with the pre-sent invention;
Fig. 9 is an axial cross-section showing a hydraulic pump; and Figs. lOa and lOb are diagrams showing the velocity triangles of the pump of the present inven-tion.
A known low cost turbine 1 as shown in Fig. 1 has a housing 2 composed of a long or regular elbow and a flared exit section and has no entrance nor exit guide vanes and only the runner blades 4 mounted on the runner hub 3 assure the flow deflection. The water flow before the entrance edge EB1 and behind the exit edge EB2 of the blades is irrotational and submitted to the potential law, characterized by cur = constant ~1 ~ 7 7~ i with r the radial distance of a water particle from the turbine shaft and CU its whirl component. With reference to Fig. 1, cul = O and CU2 is the whirl component behind the blades 4 at various cylindrical flow sections is reduced to cu5 in the draft tube, composed of the flared housing 2 and the coaxially arranged cone 9. The symbols c and cm and cmS stand for the meridional water velocities before the turbine, through the runner and in the draft tube exit section.
o The turbine 1 is mounted in a cantilever manner with shaft 5 journaled in a bearing housing 6.
The shaft 5 drives an alternator 8 by means of gears 7 or pulleys and straps.
The rotational flow in the energy recuperat-ing flow duct, called the draft tube, allows a very high diffusor angle. The radial spread of the cone 9 must sharply increase to assure an effective reduction of CU2 and to reduce friction or energy losses in the draft tube. The energy losses between the surfaces of 20 the inner cone 9 and the bell shaped turbine housing can become high, thus lowering the efficiency of a long flow guiding element. For low heads H, the energy recuperation in the draft tube from cu2 to cu5 can give satisfying results. If a certain power necessitates a high flow volume Q, the turbine 1 will be of rather large dimensions. With increase of head H, the high whirl component CU2 at the blade exit edge EB2 can only be reduced to a small cu5 by a cone 9 of severe expan-sion and axis-symmetric velocity distribution will no longer exist. When the turbine 1 operates away from the best efficiency point, at for example varying head H, the presence of a vortex core, called cork screw in addition to a whirl results in non axis-symmetric flow.
This will result in a stall and boundary layer separa-tion in the draft tube reducing furthermore its effi-ciency.
_ 7 _ 21 ~7~, Since a stall can lead to vibrations, such a cone equipped draft tube must be of a very sturdy construction. Eor this reason, cones made of concrete cannot be used, since the irregular flow will demolish surface sections and the cone is ultimately washed away.
The turbine should have by preference a ver-tical shaft. The overall efficiency 1l of such a tur-bine 1 will be rather low and depends on the reduction o of cu2 to cus An energy loss resulting from a whirl component cus at the blade 4 exit edge EB2 must be carefully weighted against the whirl component at the draft tube exit and must be added to the friction losses of the cone-equipped bell shaped draft tube.
This will determine if such a set having a turbine without guide vanes is economically viable.
The embodiment of the invention shown in Fig. 2 relates to a hydraulic turbine lO which is supported by a stationary cantilevered hollow tube 12 20 fixed to the housing plug 14 and extends in turbine casing 11. The casing ll is in form of a Y with equal cylindrical inner diameters in smooth communication.
The casing 11 includes a coaxial first tube leg llA and a second casing tube leg llB at an angle to the axis of the hollow tube 12 set in casing tube leg llA. The inlet and outlet sections are inclined at an angle towards the stationary hollow tube 12. The turbine 10 has no entrance nor exit guide vanes and has two counterrotating runners R1 and R2. Each of the rum~ers 30 R1 and R2 has at least two blades P1 and P2 of equal outside and hub diameter and a reversed pitch which is of equal or different value. The entrance runner R1 is arranged on a rotating solid shaft Sl coaxially mounted in the counterrotating hollow shaft S2 which supports the exit runner R2. The runner R2 includes a bearing 20 mounted to the shaft S1 and a bearing 21 mounted to the stationary tube 12. The shaft S2 is journaled in bearing 15 set in the housing 14. The shaft S1 is journaled in the hollow shaft S2 and supported by bearing 20 and bearing 17.
The mechanical seal assembly 19 (or other suitable seal) excludes water entering the space between the hollow rotating shaft S2 and the solid shaft S1. The mechanical seal assembly 13 (or other suitable seal) separates the turbine 10 from the ambient atmosphere. The entrance hub 22 screws onto the shaft S1 with counter threads to the rotation shaft S1 so that the hub will always be firmly tightened against spacer 25. The exit runner hub 24 is keyed to the hollow shaft S2 and tightened by an outside threaded insert ring 26 with an inside split conical bushing.
Pulley BP'R2 is threaded onto shaft S2. The shaft S2 is journaled in bearing 15 set in the housing plug 14. The shaft S1 is supported at its exit end by bearing 17 carried by pulley BP'R2. The pulley BP'R1 is directly connected to the shaft S1. Housing plug 14 is 20 firmly retained in casing tube leg llA. A machine bolt 23 is shown fastening plug 14 to the casing 11. The interior diameter of tube leg llA is at least greater than the diameter of the runners R1 and R2 so that the housing plug 14 and the turbine 10 can be removed through the opening formed at the end of casing tube leg llA.
By this arrangement, the complete turbine can be withdrawn for inspection of repair purpose or for storing during interruption of activities.
The blades Pl and P2 are made for utmost simple fabrication from flat plate material. Once traced and cut, the outside section E5 of blade P1 is twisted afterwards relative to the hub section E6 in such a way that E5 and E6 remain flat. The flat blade foot FE6 is inserted in the straight slots 16 of hub 22 of Fig. 2 and then screwed, welded or glued and after-wards machined to the outside diameter tolerances.
~7 ~7 ' Blade P2 with blade foot FE7 (not shown) is fabricated and installed in a similar manner. Fig. 2a is the view of a blade P1 in the direction of the axis a-a and Fig. 2b is the view normal to the turbine axis a-a of the blade cascade P1 and P2 unrolled in a plane.
Twisting the blades P1 and P2 must exclude a spring effect during manufacturing or under load. The blades of non rusting material should be of a short radial length to keep the twist between outside and inside sections small and thus the bending arm towards the hub is also short. For shockless entry the entrance edges of the rotating flat blades P1 and P2 are elliptically rounded off. Blunt trailing edges are not recommended and the blades should be sharpened at the exit edge E3 and E4. If special conditions require, arc curved blades of equal thickness (Fig. 4) or profiled ones must be used, but they are generally more complicated to fabricate. The entrance edges E1 and E2 of the runner blades P1 and P2 are either in the radial 20 direction or skewed at a small angle with the radius in which case the exit edges E3 and E4 will also be skewed. The rotor R1 has a rotational speed n Rl and a circumferential velocity Urunner R1, the rotor R2 has a rotational speed n R2 and the circumferential velocity Urunner R2 with n R1 and n R2 being equal or different.
The water velocity before blades P1 and behind blades P2 is c and the relative water velocities in the water channels are wp1 and wp2. The acute angle of the slots 16 and 18, (~0O-~) in Fig. 3 and the twist depend on 30 the working data H, Q and n.
In conventional one-stage axial flow turbines the guide vanes must be welded along their contour to the bearing housing and to the turbine casing and a welding gauge is required. The radial bearing must then be carefully centered and this requires much manual work.
~1 ~ 7 I
In contrast to this, the fabrication of the turbine 10 with counterrotating runners R1 and R2 is simple and no guide vanes are required.
Reference is now made to Figs. 3a and 3b. A
fluid particle on its way through the turbine runners R1 and R2 remains on the surface of a cylinder, an assumption which holds true for hydraulic efficiencies ~h close to the maximum value. A blade P1 and P2 is unrolled in a plane, as Fig. 2b shows. In the absence o of whirl CU the absolute velocity cO is then cm. This Cm is resolved due to the blade peripheral speed UrunnerRl into the relative velocity w1 at the entrance edge E1 of blade P1 and the relative velocity w2 at the runner exit edge E3 produces together with Urunner the absolute velocity C3 between the runner R1 and R2.
The relative velocity wooR1 exists for a relative flow before and behind the blade P1 and allows together with the deflection ~CU=C3U - cOU
20 the calculation of the lift coefficient ~A.
The absolute velocity C3 is resolved due to the runner R2 with UrunnerR2 into w4 at the entrance edge E4 and is then reconverted due to Ws and UrunnerR2 into C6 = cm. The velocities UrunnerR1 and UrunnerR2 need not be equal. The relative flow in a channel of runner R1 and runner R2 is wp1 and wp2.
If blade P1 is arranged in the direction of WooR1 its ~A = ~ and no power can be generated. A
physical angle of attack ~R1 between the flat blade and 30 the relative flow wooRl produces w1 and w2, ~A takes up positive values, producing the forces A, Au and AaX and the turbine generates power. For curved flat plate skeletons (Fig. 4) or profiled blades the basic relations holds also but are more complicated.
To avoid cavitation the absolute velocity Cm through the turbine will be kept small and the turbine composed of the runner R1 and the runner R2 can be placed on solid ground above the highest tail water level. The danger of an accident or an electric shock is greatly reduced but strict security measures must be respected carefully and scrupulously.
In the embodiment of Fig. 4 a blade P1 or P2 with a circular arc profile is shown in a schematic view with a radius ra. The length L (Fig. 4) of the runner blades P1, P2 is of importance. The wakes of long runner blades L dissipate more slowly than those o of shorter ones, whereas a large 1 leads to a small lift coefficient ~A. If a turbine runaway occurs, runaway speed will be rather low due to the highly disturbed flow field behind the blades P1 in which the exit counterrotating runner blades P2 wade. Since the two runners R1 and R2 are mechanically coupled to the alternator rotor the runaway speed of such turbines is inferior to that of a one stage axial flow turbine and for certain generators no mechanical braking is required and the station is shut down by stop logs or a 20 simple butterfly valve arranged at the penstock entry.
In the draft tube D of an axial flow tur-bine 10 (Fig. 2) with two counterrotating runners R1 and R2 without entrance and exit guide vanes, the kinetic energy lost is that of the water flow at the final turbine discharge cross section. This energy depends on cm2/2 and can necessitate the installation of a draft tube which will be of simple geometric con-struction, either flared or consists of a pipe with a sudden enlargement of the turbine exit diameter.
The operation of a plant consisting of a turbine 10 with two counterrotating runners R1 and R2 and its generator at variations of the head H will now be described. For load control the frequency f of the autonomous AC grid remains constant and the rotational speed of the runners R1 and R2 therefore keeps its constant value, irrespective of the operating value of H. For an actual head H the strict relation - 12 - 2i ~
llhgH = u ~CU
dictates for a blade speed u the value of ~CU and the components of the absolute velocities cO and C3 of the runners R1 and R2 (Fig. 3).
In the event the head H increases, the blade speed u is kept constant by the frequency control sys-tem. and the deflection ~CU must therefore lncrease.
Since the blade angle (Boo-~) (see Fig. 3) cannot be altered, an increase in ~CU can only be obtained by a o decrease of cm/ the meridian water velocity through the runner resulting in a reduction of the water volume flow Q. The unavoidable presence of shock losses will lower the hydraulic efficiency value of the turbine with two counterrotating runners. The power output of the turbine can therefore be inferior to the power of a similar turbine designed for this higher value H.
In the event the head H decreases, the blade speed u is kept constant by the frequency control sys-tem and the deflection ~CU must therefore decrease. A
20 decrease in ~CU can only be obtained by an increase of Cm the meridian water velocity through the runner ~Fig. 3) resulting in an increase of the water volume flow Q. It could be possible that ~CU becomes so small that the increase in cm and therefore of the volume flow Q through the rotor plate cascade suffers little deflection. A sharp decrease in efficiency will override the tendency to increase the volume flow and the turbine can produce less or no power at all.
When there are variations of the head H and 30 of frequency f with the turbine and its asynchronous generator connected to an existing AC grid having small variations of the frequency f, the grid linked fre-quency f of the asynchronous generator is within the speed droop ~p and will be for a large grid 1% to 2Q6, for a smaller one 5% or 7% or higher. The rotative velocities of the set are insignificantly modified and - 13 - 2i ~7 7~i the afore conclusions remain therefore valid. The power output will be slightly altered.
The blades P1, P2 of the turbine are posi-tioned for the rated values of the head H and volume flow Q. For a downward variation of the head H, the stilling basin must have some spill or a skip arrange-ment limiting its lowest head water level H minimum and allowing only an upwards varying head water level and the turbine layout must be made for the lowest value o of H.
In the event the rated head H of the turbine is very high, the lift coefficient ~A of blades P1 and blades P2 increases and the angle of incidellce ~ will be larger. A satisfying solution is that the runner R1 takes up a smaller head HR1 than the runner R2 which takes up a higher head HR2. This is achieved when the circumferential speed u of runner R1 is lower than the circumferential speed of runner R2. Since 71hgH = U~Cu 20 and UrunnerRl < UrUnnerR2 therefore HrunnerR1 < HrUnner R2 with HrunnerR1 + Hrunner R2 = H. Due to the different transmission ratios srunnerRl and Srunner R2 the power of runner R1 and runner R2 are cumulated in the mechanical alternator drive. The exit edge El of runner blades P1 shed a small vortex sheet and the stronger vortex sheet of runner R2 dissipates in the draft tube D without shock losses.
The rotor of the alternator A (Fig. 5) is the carrier of the magnetic field and is coupled with the turbine 10. For a small hydraulic set supplying its own autonomous grid, synchronous alternators must be used.
The alternator A is preferably single phased. The maxi-mum overspeed of the alternator A is about 1.3 of a rated alternator. An effective emergency shut down device such as a mechanical brake or a by-pass outlet - 19 ~ 7~ i can be required. Asynchronous generators can support a high runaway speed.
If the turbine is separated from its grid, the highly disturbed flow field of the two counter-rotating runners R1 and R2 will effectively limit the overspeed.
The turbine 10 can drive a low cost high rotational speed generator, i.e. alternator A or several generators of the same rotational speed.
0 The runners R1 and R2 can be coupled to the alternator A by different mechanical elements, such as bevel gears, by a gear arrangement, by a spur gear and a timing belt pulley, or by two timing belts. It is also possible to have the alternator rotor RT and the alternator casing ST directly coupled to the counter-rotating shafts S1 and S2 of the runners R1 and R2 (Fig. 8).
The greatest flexibility for coupling the turbine 10 to the alternator A is provided by the use of two timing belts (Figs. 5 to 7). On the shaft SA
belt pulley BP1 is coupled by the double gear belt BTD
with a belt pulley BP'R1 on the shaft S1. The diameter and the distance of the idler belt pulley BP3 determine the number of teeth in the mesh of the belt pulley sP1.
The rotation of the runner shaft S2 is transmitted to the alternator A by the pulley PB'R2 and the pulley BP2.
The transmission relations are either HrUnnerRl = HrunnerR2 SRRl = SRR2 HrUnnerRl + HrUnnerR2 = H
30 or HrUnnerRl i~ HrUnnerR2 SRRl ~ SRR2 HrUnnerRl + HrUnnerR2 = H
The conventional speed governor necessary for volume flow control cannot be used for the inventive turbine since the runner plates P1, P2 are of fixed geometry. For units connected to an existing grid the speed control is omitted. If the station supplies a -1S- ~ly9~
grid by itself, the frequency is kept constant by load control. For this operation the link between the tur-bine-alternator set produced power PA, the industrial load Pindustrial and the frequency control load Pfrequ is expressed by PA Pindust+Pfrequ Any difference between the available PA and the continuous variation of the industrial load Pil~dUst results in a frequency deviation tendency of the alter-o nator A. The continuously monitored alternatorfrequency is compared with the very accurate quartz reference frequency and the Pfrequ load, composed of suitable induction free resistors, is switched at a high speed across the alternator terminals. The power balance is observed and the frequency f remains con-stant. These resistors can be used for secondary indus-trial uses, such as water heating, drying, etc. This system is always dynamically stable, irrespective of the self control of various kinds of load. The total 20 dissipating capacity of the load frequency control system must be greater than the nominal power available at the alternator terminals to avoid that the frequency f goes up. If the turbine is overloaded the frequency f will go down. For a judicious partition between primary and secondary industrial load the overall operational efficiency which is of no dominant concern can nevertheless be high. The electronic security system, protecting the turbine and the alternator A keeps the various components under continuous surveillance and 30 limiting values of various parameters provoking an emergency shut-down.
Such a very sensitive frequency control sys-tem together with a safety system is patented (Canadian patent 1,244,081 - November 1, 1988 - Patent holder:
Dr.-Ing. H. Netsch).
- 16 - ~ ?~
The described turbine may be installed pref-erably for users far away from the electric public grid but can also be connected to an existing AC grid.
In summary the turbine can be equipped with a draft tube which can harness very low and very high heads with a positive suction head capable to operate at head fluctuations with a positive suction head avoiding thus turbine submergence and uneconomical civil engineering work. The turbine can be installed o with horizontal, vertical or inclined shaft and is practically maintenance free or of low maintenance. The turbine operates under very high heads for which one stage axial flow turbines cannot be developed and divides the available head potential energy evenly or unevenly between the two runners Rl and R2, so that the runner R1 at the turbine entrance produces a small power. The exit whirls of runner Rl are then of very small magnitude whereas the following counterrotating runner R2 produces a higher power with stronger exit 20 whirls dissipating in the draft tube with the absence of shock losses resulting in high hydraulic efficiencies. The power is cumulated in the mechanically coupled alternator. The turbine has a low runaway speed since the flow field of both runners R1 and R2 becomes very irregular with increased rotational speed n. In addition, the counterrotating runner R2 wades in the irregular flow field of the entrance runner R1. For certain types of generators a braking system can be omitted and falling stop logs or an~0 inexpensive butterfly valves shuts down the water flow.
The turbine is coupled by gears or a timing belt system to one or more high speed inexpensive generators and can be coupled directly to the counter-rotating stator and rotor of the generator. High fre-quency accuracy of said turbine with counterrotating runners supplying its own grid is assured by load con-trol (Canadian patent l,24~,081 - November 1, 1988 -- 17 - ~ 1 Y ~
Dr. Ing. H. Netsch), producing stable operation inde-pendent of the nature of the load. The frequency regu-lator has a satisfying degree of overall efficiency for skillful repartition of the load. The turbine generator set can be connected directly to an existing AC grid, operating thus with the frequency of the grid, dispens-ing with the frequency regulator, requiring only an emergency shut down system. Since the delivered power remains constant the hydraulic efficiency remains high, o also for frequency variations of the existing AC grid, since all flow velocities in the turbine are equally altered.
The complete cantilevered turbine 10 with two counterrotating runners R1 and R2 can be withdrawn for inspection or repair purpose or for storing during interruption of activities.
When blocking pulley BP'R1 and causing the then stationary blades P1 to become guide vanes, only the exit runner R2 rotates. The one stage cantilevered 20 turbine operating in the Y bend casing 11 can com-pletely be withdrawn.
The hydraulic engine of the present invention may also be in the form of a hydraulic pump with two counterrotating rotors without entrance and exit guide vanes. Hydraulic centrifugal or axial flow pumps should be operated at high rotational speed n r.p.m. and are coupled to high speed synchronous, or to less expen-sive, asynchronous motors of 3600 r.p.m. (3000 or other). For a certain power these 3600 r.p.m. motors are less costly and less heavy than units of lower speed and consequently the driven pumps are of smaller dimension. Centrifugal pumps have a high delivery head H and a small volume flow Q whereas axial flow pumps have a low delivery head H and a large volume flow Q.
An axial flow pump having two or more counterrotating rotors, without entrance and exit guide vanes, can have, for the same volume flow Q, a higher delivery - 18 - ~I Y~7~, head H (which a one stage axial flow pump cannot gener-ate) and has a small constructional length and is not costly to fabricate (Fig. 9).
The hydraulic axial flow pump 50 shown in Fig. 9 with two counterrotating rotors has no entrance or exit guide vanes, and has a casing 51 in the form of a Y bent with equal cylindrical inner diameters with inlet and exit sections inclined at an angle towards the stationary hollow cantilevered support tube 52 o which is fixed to the housing 54. As described in relation to Fig. 2, the Y shaped casing 51 has a tube leg 51A that is coaxial with tube 52, and a tube leg 51B at an angle to the axis of tube 52. The interior diameter of tube leg 51A is greater than the diameter of rotors PP1 and PP2 so that the plug 54 and pump 50 can be easily removed from the casing 51 through the tube leg 51A. The driven rotors PP1 and PP2 have at least two blades RP1 and RP2 of equal outside and inside diameters. Inside the hub 74 of the rotor RP2 is 20 a bearing 70 mounted to the driven shaft SP1 and a bearing 71 mounted on the stationary tube 52. The mechanical seal assembly 59 (or other suitable seal) excludes water entering the space between shafts SP1 and SP2 and the mechanical seal assembly 53 (or other suitable seal) separates the pump from the ambient atmosphere. The hollow shaft SP2 is journaled in bearing 55 set in the housing plug 54. The shaft SP1 is supported by bearing 57. The entrance hub 72 is screwed on the driven shaft SP1 which is threaded in the 30 countersense of rotation so that the hub 72 will always be firmly locked. The exit rotor hub 74 is keyed to the hollow shaft SP2 and locked by an outside threaded ring 76 having a split conical insert. The driven exit rotor RP2 is arranged on a hollow shaft SP2 of motor MP2 surrounding coaxially the shaft SP1 of motor MP1. Not excluded are driven systems consisting of spur gears, bevel gears or by two timing belts, as shown in Fig. 5 - 19 21 q',/,j j to 7. The blade PP1 of entrance rotors RP1 has an entrance edge EP1 and an exit edge EP3 and the blade PP2 of the exit rotor PR2 has the entrance edge EP2 and the exit edge EP4.
As described in relation to the turbine 10 of Fig. 2, the pump 50 can be removed with the plug 54.
Plug 54 is firmly installed in tube leg 51A by means of machine bolt 73. The plug 54 can be removed with tube 52 and pump 50 through the opening at the end of tube o leg 51A.
Thus for inspection, maintenance, repair or for operation interruption, the whole cantilevered pump can be withdrawn from its housing 51.
The blades are made for utmost simple fabrication from flat plate material and after tracing and cutting, the outside section of blade PP1 is then twisted relative to its hub section in such a way that these sections remain plane. For the blades PP2 of rotors RP2 the process is similar. The blades are made 20 from non rusting material. The foot of a blade is inserted into slots 56 or 58 and welded, screwed or glued to the respective hub 72 or 74. The blades should be short with elliptically rounded off entrance edges having no blunt trailing edges.
In Fig. lOb the velocity triangles for a radius r of the rotating and counterrotating blades PP1 and PP2 of an axial flow pump are shown. ~o entrance whirl exist, therefore CU = ~ and the absolute water flow velocity is Cpm which is resolved due to urOtorRpl 30 into the relative velocity wp1 at the entrance edge EPl of blade PP1 and the relative exit velocity wp2 at its exit edge EP3 produces the absolute velocity Cp3 between the rotor RP1 and RP2. This absolute velocity cp3 is resolved due to UrotorRP2 into wp4 and ~p5 together with urOtorRp2 produces cp6 = cpm(Cu = ~) the absolute water velocity at the exit of the pump.
Fig. lOa shows the blades, unrolled in a place similar - 20 - ~' I Y ~1 / () j to Figs. 3a and 3b. The velocity ~CpU together with WooPR1 allow the calculation of the lift coefficient ~A
and the inclination angle ~PR1 necessary to transmit the forces A, Au and AaX to the blade PP1. For generating a higher pressure head H the blades PP1 and PP2 of equal thickness are cambered in the form of a circular arc (Fig. 4) or profiled. To avoid priming of the pump the blade PP2 must be partially submerged otherwise a priming device must be provided.
o If the delivery head H is excessively high the blades PP1 and PP2 will shed wakes and the flow condition of the rotor RP2 will be disturbed. The following solution will correct this problem. The rotor RP1 generates a smaller pressure head HrotorRl whereas the rotor RP2 generates a larger pressure head HrotorR2 -delivery head HrotorRl < delivery head HrotorR2 Wakes shed by the plate exit edge EP3 will be of small magnitude and stronger wakes shed by the exit 20 edge EP4 of rotor R2 will dissipate in the flow without shock losses. The head HrotorRl and HrotorR2 will add up requiring a different power of motor MP1 and MP2. A
diffusor of suitable shape converts kinetic energy at the rotor exit to pressure.
If the volume flow Q decreases from its rated-value Q and constant speed is maintained, the head H increases and the efficiency falls off and the driving motors will be overloaded, as is also the case for single stage axial flow pumps. A stilling basin and 30 the high reservoir should be provided. Throttling valves should not be installed so that no markedly increase of the pressure head H and overloading the motor will occur. This very unfavorable operational condition which characterizes, also, one stage axial flow pumps, can be attenuated for pumps with counterrotating blade cascades if the variable speed - 21 - ~ l Y i /~ I' motors MP1 and MP2 are installed and operated in such a way that the delivered volume flow Q and the delivery head remain close to the rated design values of A
and H.
Back flow from the high to the low reservoir in case of shutdown or power failure can be avoided by a siphon in the pressure duct with a suitable aeration device.
When blocking shaft SP1, and causing the o stationary blades RP1 to become guide vanes, only the exit rotor RP2 rotates. The one stage pump operating in the Y bend casing 51 can completely be withdrawn through leg 51A.
Claims (10)
1. A hydraulic engine assembly comprising a Y
shaped casing having a pair of casing tube legs with equal inner diameters, the hydraulic engine including a cantilevered tube member enclosing at least a drive shaft coaxial with one of the casing tube legs, the casing having an inlet and the other of the tube legs being an outlet section where the outlet section is inclined at an angle towards the cantilevered hydraulic engine shaft.
shaped casing having a pair of casing tube legs with equal inner diameters, the hydraulic engine including a cantilevered tube member enclosing at least a drive shaft coaxial with one of the casing tube legs, the casing having an inlet and the other of the tube legs being an outlet section where the outlet section is inclined at an angle towards the cantilevered hydraulic engine shaft.
2. The hydraulic engine assembly as defined in claim 1, wherein the hydraulic engine includes two counterrotating entrance and exit runners of equal inner and outer diameters, the entrance runner being mounted on a solid drive shaft and the exit runner being mounted on a hollow shaft coaxially with the solid shaft, the solid and hollow shafts are supported for rotation in said cantilevered tube member.
3. The hydraulic engine assembly as defined in claim 1, wherein the hydraulic engine includes at least a runner mounted on the drive shaft and having an outer diameter wherein the outer diameter of said runner is less than the inner diameter of one of the casing tube legs.
4. A hydraulic engine assembly as defined in claim 3, wherein the hydraulic engine can be withdrawn from the casing through the opening of said one of the casing tube legs.
5. A hydraulic engine assembly as defined in claim 2, wherein the entrance runner is threaded on its solid supporting shaft with a counterthread so that the entrance runner will always be tightened on its solid supporting shaft during operation.
6. A hydraulic engine assembly as defined in claim 2, wherein the exit runner is keyed on said hollow supporting shaft and tightened by a disk threaded at its outside diameter with an inner tapered split bushing so that on tightening the set screw the exit runner will solidly press on said hollow shaft.
7. A hydraulic engine assembly as defined in claim 2 with a mechanical seal or other suitable seal element between the entrance and exit runners to prevent water or contaminants from entering bearings of the exit runner, and a second mechanical seal element is located in the exit runner to isolate the bearings from the ambient atmosphere.
8. The hydraulic engine assembly according to claim 2, wherein the engine is a turbine driving an electric generator with an input shaft and wherein the entrance runner, and the exit runner of the turbine are connected to transmission means so that the power generated by the two counterrotating runners will be cumulated at the generator input shaft.
9. The hydraulic engine assembly as defined in claim 8, wherein the head of the turbine is for reasons of operation requirements unequally divided between the counterrotating blades where these runners have the same or different rotating speed.
10. The hydraulic engine assembly as defined in claim 2, wherein the engine is a pump being driven by motor means wherein the entrance and exit runners are pump rotors rotating in opposite directions.
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| CA002199781A CA2199781A1 (en) | 1997-03-12 | 1997-03-12 | Hydraulic engines with at least two counterrotating runners |
Applications Claiming Priority (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| CA002199781A CA2199781A1 (en) | 1997-03-12 | 1997-03-12 | Hydraulic engines with at least two counterrotating runners |
Publications (1)
| Publication Number | Publication Date |
|---|---|
| CA2199781A1 true CA2199781A1 (en) | 1998-09-12 |
Family
ID=4160155
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CA002199781A Abandoned CA2199781A1 (en) | 1997-03-12 | 1997-03-12 | Hydraulic engines with at least two counterrotating runners |
Country Status (1)
| Country | Link |
|---|---|
| CA (1) | CA2199781A1 (en) |
Cited By (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| CN110046420A (en) * | 2019-04-10 | 2019-07-23 | 中国农业大学 | A method of for determining inclined shaft pump runaway speed under different leaves angle |
| CN110487551A (en) * | 2019-09-19 | 2019-11-22 | 贵州电网有限责任公司 | Hydro turbine governor servomotor is failure to actuate time emulation test system and method |
| WO2022043529A1 (en) * | 2020-08-28 | 2022-03-03 | Ocean Solution Energie | Turbopump and hydro-electric power plant including such a turbopump |
-
1997
- 1997-03-12 CA CA002199781A patent/CA2199781A1/en not_active Abandoned
Cited By (5)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| CN110046420A (en) * | 2019-04-10 | 2019-07-23 | 中国农业大学 | A method of for determining inclined shaft pump runaway speed under different leaves angle |
| CN110046420B (en) * | 2019-04-10 | 2020-09-29 | 中国农业大学 | Method for determining runaway rotating speed of inclined shaft pump under different blade angles |
| CN110487551A (en) * | 2019-09-19 | 2019-11-22 | 贵州电网有限责任公司 | Hydro turbine governor servomotor is failure to actuate time emulation test system and method |
| WO2022043529A1 (en) * | 2020-08-28 | 2022-03-03 | Ocean Solution Energie | Turbopump and hydro-electric power plant including such a turbopump |
| FR3113706A1 (en) * | 2020-08-28 | 2022-03-04 | Ocean Solution Energie | TURBOPUMP AND HYDRO-ELECTRIC POWER PLANT INCLUDING SUCH TURBOPUMP |
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| EEER | Examination request | ||
| FZDE | Dead |