Note: Descriptions are shown in the official language in which they were submitted.
' 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.
- 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
- 3 -
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
- 10 -
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.
