Note: Descriptions are shown in the official language in which they were submitted.
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Description
A Dual Pressure Turbine
Technical Field
This invention relates generally to a heak
recovery system adapted to use the heat produced by
the engine to generate a gas t:hat is subsequently used
to drive a dual pressure turbine for producing useful
work and more particularly to the components of the
system and their interrelationship coupled with the
structure of the dual pressure turbine.
_ack~round Art
Heat recovery systems are provided in a
system to take full advantage of the total energy
being produced by an engine. A large amount of energy
produced in an engine is lost through the exhaust
system and jacket water cooling. Many of the known
heat recovery systems use various forms of heat
exchangers in the exhaust system to convert the heat
in the exhaust to a form that subsequently performs
useful work. The steam-based Rankine bottoming cycl~
heat recovery principle is well known and has the
potential to increase total engine performance by
utilizing the engine exhaust to perform useful work.
One process frequently used converts water to steam
and uses the steam to operate miscellaneous services,
such as heaters, and to drive a steam turbine. One of
the major problems encountered when using the engine
exhaust system to convert water to steam is soot
fouling of the heat exchanger or boiler that is
located in the exhaust system. This soot fouling
problem is even more pronounced on systems used in
diesel engines. It has been found that the soot
thickness is strongly dependent on the temperature of
the walls of the tubes in the boiler
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that are located in the exhaust system. Naturally the
more tubes that are used coupled with the lower
temperature used in the tubes, the more collection of
soot. To optimize a system to accept a 5% loss due to
sooting could add 30% to the size and cost of a total
system.
Another major problem that is associated
with the low tube temperature in the boilers that are
located in the exhaust system is the tendency for
various gases to precipitate out of the exhausk gas
due to the lower tube temperature. These gases form
oxides that chemically attack the various metals in
the exhaust system thus shortening their useful life.
Various forms of turbines have been used in
the Rankine bottoming cycle systems. These range from
single or multiple stage high pressure turbines to
single or multiple stage low pressure tur~ines and to
a mix of low pressure and high pressure turbines. The
main objectives considered in determining the type of
turbine to use is maximizing performance, controlling
the cost versus performance, and controlling total
package size. It has been generally ~ound that the
more stages used in the turbine, the greater the
system efficiency. However, both the cost and si~e of
the turbine increases with an increase in the number
of stages used. Many times the cost of added stages
will increase at a much higher rate than that of the
system performance.
Of thel~ual pressure turbines used, some of
them direct the high pressure steam to only the high
pressure side and the low pressure steam to the low
pressure side and subsequently to the outlet port.
These normally fail to efficiently use all of the
available work in the steam or they have to use many
turbine stages thus adding significantly to the total
cost. Of the other dual pr~ssure turbines used, the
exhaust steam from the high pressure stage or stages
is directed to a mixing chamber where it mixes with a
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low pressure steam and directed into the low pressure
stages. ~he mixing chamber is primarily provided to
ensure that the temperature, pressure, and velocity of
the steam from the high pressure stage is respectively
equalized wit the temperaturel pressure, and velocity
of the low pressure steam prior to entering the low
pressure stages. The mixing chamber that is normally
used increases the size of the total package further
adds the possibility of unwanted turbulence in the
chamber.
The dual pressure turbine may be of a kind
which includes a housing having a first stage disposed
therein, the first stage including a rotor having a
plurality of blades secured around the periphery and
being rotatably positioned in a rotor chamber, the
high temperature and high pressure inlet port, an
outlet passage connected to the rotor chamber and a
plurality of nozzles interconnecting the high
temperature and high pressure inlet pork with the
rotor chamber adjacent a preselected number of blades
of the plurality of blades, the plurality of nozzles
being adapted to increase the velocity of the high
temperature and hlgh pressure gas so that the gas
impinges on the plurality of blades at a supersonic
velocity; and may further include a second stage
disposed in the housing, the second stage including a
second rotor having a second plurality of blades
secured around the periphery and being rotatably
positioned in a,second rotor chamber, the low pressure
and low temperature inlet port, and a second plurality
of nozzles opening into the second rotor chamber
adjacent the plurality of blades and being evenly
spaced therearound
The present invention is directed to
overcoming one or more of the problems as set forth
above.
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Disclosure of the Invention
In one aspect of the present invention, a
dual pressure turbine is provided to generate an
output force in response to the flow of a gas
therethrouyh. The dual pressure turbine includes a
housing having a first inlst port adapted for
connection with a source of gas at a preselected high
pressure level and superheated to a preselected high
temperature. A first stage is defined in the housing
and includes a first rotor chamber, a first rotor
having a plurality of blades secured around the
periphery and is rotatably positioned in the first
rotor chamber, a first plurality of nozzles
interconnecting the first inlet port and the first
rotor chamber adjacent a preselected number of blades
of the first plurality of blades, and an outlet
passage connected with the first rotor chamber
adjacent at least the preselected number of blades. A
second stage is defined in the housing and includes a
second rotor chamber, a second rotor having a second
plurality of blades secured around the periphery and
is rotatably positioned in the second rotor chamber
and drivingly connected to the first rotor, a second
plurality of nozzles opening into the second rotor
chamber adjacent the second plurality of blades and
the nozzles are evenly spaced therearound, and an
outlet passage connected to the second rotor chamber.
A second inlet port is located in the housing and is
adaptad for connection with a second source of gas at
a preselected pressure lower than the preselected high
pressure and sup~rheat~d to a temperature lower than
the preselected high temperature. A first passage
means is provided for connecting the second inlet port
with the second plurality of nozzles so that the gas
from the second source of gas is substantially a~ially
directed to the second plurality of nozzles at a
predetermined velocity and a second passage means
connects the outlet passage of the first stage with
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the second plurality of nozzles so that the gas from
the outlet passage of the first stage is substantially
axially and substantially separately directed to the
second plurality of nozzles at substantially the same
velocity as the predetermined velocity of the gas from
the second source.
The present invention provides a heat
recovery system that fully utilizes the heat generated
by the engine to drive a dual pressure turbine to
produce useful work. The coolirlg system or more
specifically the jacket water of the engine absorbs
sufficient heat energy and the jacket water is passed
through a heat exchanger to convert a fluid to a gas
and then superheat the gas for effective use in the
turbine. Another heat source from the engine/ such as
the oil lubrication system, provides sufficient heat
to preheat the fluid b~fore it enters the third heat
exchanger located in heat transferring relation to the
exhaust system. This relationship maintains an
increased temperature at the inlet port of the heat
exchanger thus reducing the sooting problem and the
precipitation of acidic gases. The first and second
passage means in the dual pressure turbine cooperate
to ensure that the temperature, pressure, and velocity
of the respective gases from the outlet passage of the
first stage and the second source are respectively
substantially equal before entering the second stage.
This arrangement eliminates a mixing chamber between
the first and second stages thus also eliminating the
possibility of turbulence that could exist in the
mixing chamber.
Brief Description of the Drawings
Fig. l is a partial schematic and
diagrammatic representation of a heat recovery system
incorporating an embodiment of the present invention;
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Fig. 2 is a partial sectional view,
illustrating in greater detail, one of the components
shown in Fig. l;
Fig. 3 is a partial section taken along the
line III-III of Fig. 2;
Fig. 4 is an enlarged partial section of one
of thQ nozzles and three of the rotor blades of the
first rotor of FigO 2 shown in their true position;
Fig. 5 is a partial section taken along the
line V-V of Fig. 2; and
Fig. 6 is a section taken along line VI-VI
of Fig. 5.
Best Mode for Carrying Out the Invention
Referring now to the drawings, and more
specifically to Fiy. l, a heat recovery system 10 is
provided for use on an engine 12 haviny a liquid
cooling system 14 and an exhaust system 16. The
liquid cooling system 14 includes a pump 18 for
circulating cooling fluid through a known water jacket
(not shown) in the engine, a thermostat 20 adapted to
control the temperature of the coolin~ fluid, and the
associated cooling lines 22. The exhaust system 16
includes an exhaust manifold pipe 24 connected to the
25 engine 12.
The engine 12 further includes an oil
cooling system 26. The oil cooling system 26 includes
an oil pump 28 for circulatin~ the hot lubricating oil
of the engine 12 to an oil cooler 30 and the
connecting lines 32 and 34 to provide a path for the
continuous circulation of the oil from the engine 12
to the oil cooler 30 and back to the engine 12.
The system 10 includes a source of
pressurized fluidr such as a pump 40. The pump 40
draws a fluid, such as water, from a reservoir 42 by
way of a conduit 44. A fluid polisher 46 is provided
in line 44 and is adapted to remove minerals and other
deposits from the water. A source 48 of make-up water
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is provided to replenish the water in the reservoir 42
if it becomes low.
A means 50 is provided for preheating the
water from the pump 40. The preheating means 50
includes the oil cooling systlem 26 and a housing 52
having a first fluid passagew,ay 54 connected to the
pump 40 by an conduit 56~ A relief valve 57 is
connected to the conduit 56 and controls the maximum
pressure of the fluid in the conduit 56. The housing
52 is located in the conduit 32 of the oil cooling
system 26 and defines a second fluid passageway 58
interconnecting the conduit 32 within the housing 52.
The second fluid passageway is operatively associated
in heat transferring relation with the first fluid
passageway and is adapted to transfer the heat energy
in the hot oil from the engine 12 to the water passing
through the first fluid passageway 54.
A first heat exchanger 60 is connected to
the preheating means 50 by a conduit 62. The first
heat exchanger includes an inlet port 64, and outlet
port 66, and an evaporator 68. The evaporator has a
first fluid passageway 70 connected between the inlet
port 64 and the outlet port 66. A second fluid
passageway 72 is located in the evaporator 68 and
connected to the cooling lines 22 so that the pump 18
continuously circulates the engine cooling fluid. The
second fluid passageway 72 is operatively associated
in heat transferring relation with the first fluid
passageway 70 s~ that the heat energy in the engine
cooling fluid in the second fluid passageway 72 is
transferred to the fluid or water in the first fluid
passageway 70. This heat energy is sufficiently high
to convert the water in the first fluid passageway 70
to a steam.
A second heat exchanger 74 has an inlet port
76 and an outlet port 78 and located in the exhaust
manifold pipe 24. A conduit 80 connects the inlet
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port of the second heat exchanger 74 with the outlet
port 66 of the first heat exchanger 60.
A third heat exchanger 82 has an inlet port
84 and an outlet port 86 and is also located in the
exhaust manifold pipe 24. A conduit 88 connects the
inlet port 84 of the third heiat exchange 8~ with the
conduit 62.
A first means 90 is provided for
controllably directing a portion of the preheated
water at a preselected pressure to the inlet port 64
of the first heat exchanger 60 so that the preheated
fluid is converted to a steam or gas at substantially
the preselected,pressure and for subsequently
directing the steam from the outlet port 66 of the
first heat exchanger 60 to the inlet port 76 of the
second heat exchanger 74 so that the steam at the
preselected pressure is superheated to a pre elected
temperature. The first means 90 for controllably
directing includes a flow control valve 92 operatively
connected in the conduit 62 between the inlet port 64
of the first heat exchanger and the point of
connection of the conduit 88 with the conduit 62.
A second means 94 is provided for
controllably directing the other portion of the
preheated water, at a preselected pressure higher than
the preselected pressure of the preheated water from
the first means 90, to the inlet port 84 of the third
heat exchanger 82 so that the other portion of the
preheated fluid~is converted to a steam at the higher
preselected pressure and superheated to a preselected
temperature higher than the temperature of the staam
from the first means 90. The second means for
controllably dir~cting includes a second flow control
valve 96 located in the conduit 8~.
A dual pressure turbine 100 is provided in
the heat recovery system 10 and includes a housing 101
having a first inlet port 102 connected to the outlet
por-t 86 of the third heat exchanger 82 by a conduit
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104, a second inlet port 106 connected to the outlet
port 78 of` the second heat exchanger by a conduit 108,
and an outlet port 110 connected to a condenser 112 by
a conduit 114. The condensed fluid from the condenser
112 returns directly to the reservoir 42 for reuse in
the heat recovery syskem 10.
An output shaft 116 connects the dual
pressure turbine 100 to a common load 118 that is also
driven by the engine 12.
lG As is best shown in Fig. 2, a first stage
120 is located in the housinq 101 and includes a first
rotor chamber 122, a first rotor 124 rotatably
positioned in the first rotor chamber 122, a first
plurality of blades 126 evenly spaced and secured
around the outer periphery of the rotor 124, a first
plurality of nozzles 128 interconnecting the first
inlet port 102 and the first rotor chamber 122, and an
outlet passage 130.
For ease of illustration, a nozzle 132 is
shown in Fig. 2 rotated 90 out of location. The
nozzle 132 as illustrated should be rotated 90 to be
positioned in the top plane as properly illustratad in
Fig. 4. Each nozzle 132 of the plurality of no~zles
128, as better illustrated in Fig. ~ has a bellmouth
inlet 134 formed by a radius of curvature equal to 1/2
the diameter of a throat portion 136. The throat
portion 136 has an axial straight length equal to 1/2
the diameter of the throat. An outlet portion 138 has
a conical surfade with an included angle of
approximately 10. As further illustrated in Figs. 2
and 3, the outlet portion 138 of each nozzle 132 opens
into the first rotor ohamber 122. The plurality of
nozzles 128 cover an arc of approximately 136 and
open into the first rotor chamber 122 adjacent the
plurality of blades 126.
Each blade 140 of the plurality of blades
126 are shaped such that a passage 142 between each
pair of blades is of a constant width. The height and
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width of each blade 1~0 is substantially the same as
the height of the minor diameter of each of the outlet
portions 138 of the plurality of nozzles 128.
A second stage 150 is located in the housing
101 and includes a second rotor chamber 152, a second
rotor 15~ rotatably positioned in the second rotor
chamber 152, a second plurality of blades 156 evenly
spaced and secured around the periphery of the second
rotor 154, a second plurality of evenly spaced nozzles
158 opening into the second rotor chamber 152 adjacent
the second plurality of blades, and an outlet passage
160.
A first passage means 162 is provided for
connecting the second inlet port 10~ with the second
plurality of nozzles 158 so that the steam from a
source, such as the second heat exchangar 74, is
substantially axially directed to a first portion of
the nozzles of the second plurality of nozzles 158 at
a predetermined velocity. The first passage means
162, as better illustrated in Figs. 5 and 6, includes
an arcuate slot 164 connected to the second inlet port
106. At the area of connection with the second inlet
port 106, the arcuate slot 164 has a predetermined
cross-sectional area and defines an arc of
approximately 200. The cross-sectional area
decreases as the arcuate slot 164 extends towards the
second plurality of nozzles 158. The 200 angle of
arc remains substantially constant and the arcuate
slot opens to the second plurality of nozzles 158 in a
substantially axial direction.
A second passage means 170 is provided for
connecting the outlet passage 13Q of the first stage
120 with the other portion of nozzles of the second
plurality of nozzles 158 so that the steam from the
outlet passage 130 of the first stage 120 is
substantially axially directed to the other portion of
the pluralit:y of nozzles 158 at substantially the same
velocity as the predetermined velocity of the steam
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from the second inlet port 106. The second passage
means 170, as better illustrated in Figs. 5 and 6,
includes an arcuate slot 172 connected to the outlet
passage 130 of the first stage 120. At the area of
connection with the outlet passage 130, the arcuate
slot 172 has a predetermined cross-sectional area and
defines an arc of approximately 160. The
cross-sectional area increases as the arcuate slot 172
extends towards the second pl~lrality of nozzles 158.
The 160 angle of arc remains substantially constant
and the arcuate slot 172 extends towards the second
plurality of nozzles 158 in a substantially axial
direction.
A third stage 180 is located in the housing
101 and includes a third rotor chamber 182, a third
rotor 184 rotatably positioned in the third rotor
chamber 182, a third plurality of blades 186 evenly
spaced and secured around the periphery of the third
rotor, a third plurality of evenly spaced nozzles 188
connected to the outlet passage 160 of the s cond
stage 150 and opening into the third rotor chamber 182
adjacent the third plurality of blades, and an outlet
passage 190 connected to the outlet port 110.
It is recognized that various types of
fluids could be used in this heat recovery system
without departing from the essence of the invention,
but preferably water is used due to its availabllity
and ability to absorb large amounts of heat energy.
Industrial Applicability
During operation of the heat recovery system
10, water from the reservoir 42 is drawn into the pump
40 and directed into the conduit 56 at a predetermined
pressure level of for example 21~0 kPa (310 psi) as
established :by the relief valve 57. The polisher 4~
sufficiently cleans the water so that no impurities or
other deposits are allowad to pass throughO These
deposits, if allowed to pass through would possibly
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cause lime or other deposits to form in the system
thus reducing the overall efficiency of the heat
recovery system.
The pressurized water in conduit 56 flows
through the ~irst passageway 54 of the prehea-ting
means 50 and is preheated by the hot oil in the
lubrication system. Due to the temperature of the
oil, the water in the first passageway 54 is preheated
to approximately 105 C (220 F). Because the water in
the conduit 56 is under pressure, it does not convert
to steam.
The preheated water is directed to the
evaporator 68 of the first heat exchanger 60 by the
conduit 62. The flow control valve 92 in the conduit
62 controls the flow entering the first heat exchanger
60 and also controls the pressure of the fluid to
approximately 130 kPa (19 psi). As the preheated
water passes through the first fluid passageway 70 of
the evaporator 68, the heat in the second fluid
passageway 72 ~rom the jacket water of the engine 12
adds heat energy to the water and converts the water
to steam. Since the temperature of tha fluid in the
water jacket is approximately 121C (250F) and the
pressure of the fluid is approximately 130 kPa, the
water is converted to steam and exits at 121C and 130
kPa. The steam from the first heat exchanger 60 is
directed to the second heat exchanger 74 where it is
superheated to a temperature of approximately 182C
(360F).
The exhaust from the engine 12 is used to
superheat the steam in the second heat exchanger 74.
The temperature of the exhaust as it leaves the engine
12 is approximately 460 C (860 F). By sizing the
second heat exchanger 74, the steam is superheated to
a predetermined temperature and maintained at the
predetermined pressure level of 130 kPa.
The other portion of the preheated water in
conduit 62 is directed to the third heat exchanger 82
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through the conduit 88. The flow control valve 96 in
the conduit controls the flow entering the third heat
exchanger 82 and also controls the pressure of the
fluid to approximately 1724 ~Pa (250 psi). By
properly si~ing the third heat exchanger 82, the
preheated water entering the third heat exchanger 82
is converted to steam and superheated to approximately
415C (780F) and maintained at the 1724 kPa pressure
level.
A*ter passing across both the second and
third heat exchangers 7~,82, sufficient heat energy in
the e~haust has been used so that the temperature of
the exhaust as it leaves the exhaust manifold pipe is
approximately 204C (400F).
By using the heat energy in the oil cooling
system 26 to preheat the water from the pump 40 and by
using the heat energy in the jacket water of the
engine 72 to convert some of the preheated water to
steam, then the heat energy in the exhaust system is
sufficient to convert the remainder of the water to
steam and to superheat both quantities of steam to
their respective preselected temperatures. This
arrangement provides the greatest effective use of the
potential heat energy from the engine 12.
Furthermore, since the heat energy from the jacket
water is being dissipated, a radiator normally will
not be required in the system. By ensuring thak the
temperature of all fluids entering the second and
third heat exchangers 74,82 are above a preselected
level r the pro~lems of sooting and oxide formations
are greatly reduced. This helps both the efficiency
o~ the system and the life of the elements of the
second and third heat exchangers 74,82 located in the
exhaust system 16.
The dual pressure turbine 100 is used to
convert the superheated steam into useful work. The
superheated steam at 460C and 1724 kPa is connected
to the first. inlet port 102 of the dual pressure
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turbine 100 and the superheated steam at 121C and 130
kPa is connected to the second inlet port 106.
The superheated steam in the first inlet
port 102 is directed into the first rotor chamber 122
throuyh the first plurality oE nozzles 128. Based on
the volume of superheated steam produced and the
pressure thereof, nine nozzles 132 are used to direct
the superheated steam to the blades 126 of the first
stage rotor 12~.
Due to the shape of each oE the nine
nozzles, the velocity of the steam is increased such
that the steam impinges the blades at supersonic
speed. It has been found that the highest efficiency
of the first stage is attributed to the high velocity
of the superheated steam, the very close spacin~ of
the nozzle exits, and the controlled constant spacing
between the blades 126 on the rotor 124. Furthermore,
the height and width of each of the blades 126 is
substantially the same as the height of the minor
diameter of each of the nozzle exits 138. These
relationships provide a desired speed of 50,000 rpm.
The superheated steam leaves the blades 126
first stage rotor 124 at a velocity of approximately
137m/sec ~450 ft/sec), a temperature of approximately
182C (360F); and a pressure of approximately 130 kPa
(l9 psi). The pressure differential between the
pressure of the steam entering the first stage 120 and
the pressure of the steam leavin~f the first stage 120
results in a first stage turbine having a pressure
ratio of approximately 13:1. The superheated steam
enters the arcuate slot 172 and is directed towards
the second plurali~y of nozzles 158. Due to the
increasing cross-sectional area of the arcuate slot,
the velocity of the superheated steam is decreased to
approximately 76m/sec (250 ft/sec) prior to entering
the second plurality of nozzles 158. Because the
arcuate angle of the arcuate slot 172 is approximately
160, the superheated steam from the first stage is
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directed substantially only to a portion of the second
plurality of nozzles 158 equal to the arcuate angle of
the arcuate slot 172.
The superheated steam at the second inlet
port 106 is communicated with the second plurality of
nozzles 158 through the arcuate slot 164. The
velocity of the superheated steam at the second inlet
port is approximately 30m/sec (100 ft/sec) and has a
temperature of approximately :182C. Since the
cross-sectional area of the arcuate slot 164 decreases
in size as it extends towards the second plurality of
nozzles 158, the velocity of the steam is increased to
approximately 7ffm/sec (250 ft/sec) prior to entering
the second plurality of nozzles 158. Because the
arcuate angle of the arcuate slot 164 is approximately
200, the suparheated steam from the second inlet port
is direct~d substantially only to the other portion of
the plurality of nozzles 158 equal to the arcuate
angle of the arcuate slot 164.
It is desirable that the superheated steam
be directed into the second plurality o~ nozzles 158
in a substantially axial direction in order to
generally eliminate turbulence in the steam flow.
Furthermore, in order to maximize the efficiency of
the turbine 100, the respective temperature, pressure,
and velocity of the superheated steam from the second
inlet port 106 and the outlet passage 130 of the first
stage 120 should be substantially equal prior to
entering the second plurality of nozzles 158.
Wherein the first stage 120 is only a
partial ad~ission stage of approximately 37.8%
admission, the second stage 150 is a full admission
stage receiving the superheated steam from two
different sources. The superheated steam exiting the
second stage. 150 enters the third plurality of no~zles
188 of the third full admission stage 180. The
superheated steam exiting the third stage 180 is
connected to the outlet port 110 and directed to the
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condenser 112 where it is condensed back to water and
returned to the reservoir 42.
The heat recovery system 10 as described
above provides a system having an operating efficiency
above 20% and the dual pressure turbine 100 in the
system has an operating efficiency above 80% and most
effectively utilizes the heat energy produced by the
engine 12 while limiting total si~e and controlling
the cost. The three stage dual pressure turbine 100
effectively utilizss the superheated steam produced
from the heat energy in the engine 12 by using a first
impulse stage having steam at supersonic velocity
impinging the blades of the first stage rotor.
Furthermore the velocity of the steam exiting the
first stage and the steam entering the second inlat
port is controlled so that when they enter the second
stage their velocities are substantially the same.
The second and third stages axe full admission
transonic reaction stages. The above three stage dual
pressure turbine arrangement maintains the superheated
steam in a superheated condition all the way through
each stage. If any droplets of ~ater form while in
any of the three stages, they would erode away
material thus causing damage and shortenin~ the
effective life of the dual pressure turbine.
Other aspects, objects and advantages of
this invention can be obtained from a study of the
drawings, the disclosure and the appended claims.
