JPH0776535B2 - Dual-introduction Radial T-Vine Gas Generator - Google Patents

Description

The present invention relates to a high-performance gas turbine gas generator using a centrifugal compressor and a radiant-flow turbine, which shows high efficiency in terms of fuel consumption rate.

For example, gas turbine power plants have proven to be advantageous as an alternative to diesel engines for use in vehicles and wherever low fuel consumption small and lightweight engines are required. In actual experiments, it has been found that using a normal gas turbine is advantageous in terms of saving space and weight. However, the deterioration of the fuel consumption rate of a simple cycle engine when the rated power is reduced has been a major drawback of prior art devices.

Known gas turbine auxiliary power plants have a double-introduction low-pressure centrifugal compression stage and a single-induction high-pressure centrifugal compression stage, both compression stages being used for combustion with fuel in the combustor and for outflow for external use. Connect in series to provide compressed air to both as air. The hot gas from the combustor is then
Guided to a single stage radiant inflow turbine, it drives both compressors and is the shaft power takeoff for external use.

The inability to form a sufficient pressure ratio in the past has resulted in a much higher fuel consumption rate than high speed diesel engines. To overcome these pressure ratio limitations, the industry has one or more radiant flow compression stages upstream of the final centrifugal compression stage, or two with a conventional radiant turbine for the gas generator portion of the engine. Two single-introduction centrifuge stages were used. State-of-the-art 10 with any of these shapes
For 00 gas turbines, the industry currently expects a total pressure ratio of about 15: 1 or less and a fuel consumption rate of about 0.45.

In view of the above drawbacks of conventional gas turbine gas generator systems, it is an object of the present invention to provide a high pressure ratio gas turbine gas generator that uses a radiant inlet turbine to drive the centrifugal compressor components.

Single cycle structure (ie recuperator / rege
It is an object of the present invention to provide a gas turbine gas generator which maintains its fuel consumption) and has a fuel consumption rate comparable to that of a diesel engine.

It is a further object of the present invention to provide a high pressure ratio gas turbine gas generator using only two compression stages, which is particularly suitable for relatively low power (typically 400 or less) applications.

The aforementioned deterioration of the fuel consumption rate of the prior art small gas turbine engine has been improved by the present invention described in more detail below, so that even a small turbine engine can have high engine efficiency and thus simple cycle gas. Even though it is a turbine engine, it is possible to compete with a high-speed diesel engine even if the fuel consumption rate falls below 1000.

The present invention includes a gas turbine gas generator flow passage shape having the following characteristics as a component of the air flow (see FIG. 1): 1. A dual-introduction centrifugal compressor of basically radiant flow as the first stage, 2. Single-stage centrifugal compressor as two stages, 3. Radiant inflow turbine as gas generator turbine.

The cycle according to the invention has a particularly low power range, for example 4000
When used in the following engines, sufficiently higher pressure ratios are available than in ordinary gas turbine engines. Since the pressure ratio is one of the important factors contributing to efficiency, the cycle according to the invention produces higher thermal efficiency than existing turbine engines in the low power range.

Another important feature of the gas generator according to the present invention is the first
It is the "matching" of the first compressor stage and the second compressor stage in part by the specific speed by carefully adjusting the pressure ratio between the stage and the second stage. The radiant inflow turbine, the first stage compressor and the second stage compressor all rotate at the same angular velocity. Such a "single spool (Single spoo
l) ”Direct drive of the array compressor is at least 3
Brings one unique advantage. First, the mechanical strength of the radiant inlet turbine is sufficient to allow operation of each centrifugal compressor near its optimum level for specific speed. Second, the radiant inlet turbine operating under these conditions will operate near its optimum specific speed. Third, the high peripheral velocity as a result of the radial inflow turbine lowers the stagnation temperature of the hot compressed gas impinging on the turbine blades. As a result, the turbine inlet temperature can be increased to further increase thermal efficiency, or the operating life of turbine components can be extended.

As described herein, a compact, high efficiency gas turbine gas generator according to the present invention includes a compressor arrangement for producing a total compression ratio of greater than about 15: 1. The compressor device comprises a first stage dual induction centrifugal air compressor having a pair of inlets and a common outlet; a flow is connected to the first stage common outlet adjacent to the first stage compressor 1 A second stage single inlet centrifugal air compressor having one inlet and also a second stage outlet;
A shaft assembly for mechanically connecting the first and second stages to rotate at the same angular velocity. The gas generator includes a combustion means coupled to the second stage outlet and for combusting fuel with compressed air to generate combustion gases. Further, the gas generator utilizes a single stage radiant flow turbine with one inlet and one outlet. The turbine is operatively connected directly to the shaft assembly and combustion gases drive the first compression stage and the second compression stage. The exhaust means creates more work from the partially expanded combustion gas (work
-Producing) is connected to the turbine outlet. Importantly, the pressure ratio of the first compression stage is greater than about twice the pressure ratio of the second compression stage; the inlet Mach number of the first stage dual-introduction compressor is greater than about 1.4; The specific speeds of the two compression stages approach their optimum values, respectively, and are respectively greater than about 0.60.

Preferably, the gas generator total pressure ratio over the first compression stage and the second compression stage is about 20: 1 or greater, and the first compression stage pressure ratio is
It is between 6: 1 and 9: 1, while the second compression stage pressure ratio is between about 2: 1 and about 4: 1.

It is preferable that the specific speed of each compression stage is about 0.65 to 0.80, the specific speed of the turbine component is about 0.5 to 0.75 at the same time, and the efficiencies of both the compressor and the turbine are close to their peak values.

Almost all of the compressed air exiting the first stage during steady state operation is admitted to the second stage compressor and almost all of the compressed air reaching the second stage compressor is admitted by the combustor.

As described herein, a single spool two stage high performance compressor unit of the present invention comprises a dual introduction centrifugal compressor first stage; and a second stage compressor for receiving gas exiting the first stage compressor. A second stage operatively connected single induction centrifugal compressor; and a shaft assembly for coaxially mounting both the first and second stage compressors for dependent rotation at the same speed. . Total pressure ratio across the compressor unit is about 15: 1
Above, the pressure ratio of the first compression stage is more than twice the pressure ratio of the second compression stage, and the comparative rotation speed of each of the first compression stage and the second stage compressor is about 6.0 or more. .

Preferably, the compressor unit further comprises a single stage radiant inlet turbine mounted on a shaft assembly for driving both the first and second stage compressors at the same speed, the specific speed of the turbine being between about 0.50 and 0.75. Is.

The accompanying drawings, which are related to the present specification and constitute a part of the present invention, illustrate one embodiment of the present invention and together with the description serve to explain the principle of the present invention.

FIG. 1 is a schematic cross-sectional view of a gas turbine gas generator according to the present invention, and FIGS. 2A and 2B use a compressor component and a turbine component matching the specific speed in the gas generator made according to the present invention. 6 is a graph showing the improvement in efficiency obtained by

The details will be described with reference to the preferred embodiments of the present invention shown in the accompanying drawings.

As illustrated in FIG. 1, the gas generator 10 according to the present invention has a dual inlet centrifugal low pressure first stage having a pair of inlets and a single outlet to create a total pressure ratio of about 15: 1 or greater. Includes a compressor. The preferred embodiment compressor of the present invention includes a low pressure first compression stage 14. The first compression stage 14 has a compressor rotor 18 rotatably mounted by a shaft assembly 20 and has dual axial compression passages having two axially direct flow paths indicated by arrows 22 and 24 respectively. Machine module 16 is included. The first compression stage 14 also includes an enclosure housing 26 defining a pair of axially opposed flow symmetrical inlets 28, 30 for directing air to the compressor blade assemblies 32, 34 of the rotor 18. An improved dual induction compressor particularly suitable for this application is described in US Pat. No. 4,530,639.

The dual-introduction compressor module 16 is a diffuser assembly
A single annular, radially oriented compressor outlet 36 is operably connected to 38. The diffuser assembly 38 receives high velocity air from the outlet 36 and ultimately converts the high velocity air to high pressure, low velocity air for transmission to the high pressure compression stage 40 and to the combustor 60. The diffuser assembly 38 can be replaced with a manifold assembly (not shown) designed to hold a portion of the dynamic head of air leaving the outlet 36. Various configurations of diffuser devices that can be used as diffuser assembly 38 are disclosed in U.S. Pat. No. 4,573,868.

With continued reference to FIG. 1, the diffuser assembly 38 is an annular air plenum 42 for collecting the air being dissipated.
It is equipped with. In the embodiment shown in FIG. 1, the air reservoir 42 is connected to a cross duct 44 which is connected to an inlet air reservoir 46 of the high pressure compression stage 40, which will be described in more detail below.

Also, the compressor device according to the present invention comprises a high pressure centrifugal compressor, which is disposed adjacent to the first stage compressor and has a single inlet and a single outlet.
Including steps. As shown in the embodiment of FIG.
Further includes a high pressure compression stage 40 which is a single inlet radiative compressor with a housing 48 forming a compressor inlet 50 for receiving compressed air from an inlet air reservoir 46. The high pressure compressor housing 48 also defines a second stage compressor outlet 52 which is connected to a combustor feed air sump 54 via a second stage diffuser assembly 56. High pressure compressor rotor 58 is housing 48
Located within and pivotally mounted within shaft assembly 20. Thus, the compressed air flow path is from the collection air reservoir 42 of the double-introduction compressor module 16 to the high pressure compression stage inlet air reservoir 46 via the cross duct 44, and to the combustor supply air reservoir 54 past the high pressure compressor roller 58. move on.

Further according to the invention, the pressure ratio between the first compression stage and the second compression stage is such that the pressure ratio of the first compression stage is about 2 times the pressure ratio of the second high pressure stage.
Selected to be more than double. Further, the flow path dimensions of the first compression stage are selected to produce a preferred specific velocity corresponding to the pressure ratio selected for that stage, as will be described in more detail below. As shown in the embodiment, the total compression ratio over both the first compression stage and the second compression stage is about the pressure ratio of the first compression stage 14.
6: 1 to 9: 1 and greater than about 15: 1 at a pressure ratio of the second compression stage 40 of about 2: 1 to about 4: 1. Relatively low second
The stage pressure ratio is such that the second stage specific speed (see description below) is kept as high as possible.

A typical inducer tip relative inlet Mach number at the first stage compressor inlets 28, 30 is about 1.4 or more and 1.0 or more occurs at the outer tips of the leading edges of the blades 32, 34 during operation at rated power. The Mach number causes shocks at the compressor inlet, but these are relatively weak oblique shocks and do not seriously affect overall performance. The present invention has a design philosophy in which control of the specific speed of the compressor components is important by maintaining a low inlet Mach number. This is different from the conventional conventional design procedure.

The component efficiencies of compressors and turbines are closely dependent on choosing the best so-called specific speed. This relates to the speed-flow-work that forms the most efficient energy within its particular component. Specific speed (Ns) is Where ω = rotational speed radian / sec Q = volume flow M ³ / sec ΔH = specific output Watt / kg / S.

It has been found that the pressure ratio division between the first and second compression stages is important in order to meet the requirements for the preferred specific speed of both compression stages at the high pressure ratio of the present invention. Moreover, it is highly advantageous to use a radiant inlet turbine component where the total compressor performance should be operated near the optimum specific speed.
This matching is achieved despite the use of a single spool arrangement, with compression stages 14 and 40 directly
Driven by. These conditions are achieved with a high efficiency, high performance single cycle gas generator such as the gas generator 10 shown in FIG.

Further in accordance with the invention, the gas generator includes combustion means operatively connected to the outlet of the second compression stage and for receiving combustion fuel using compressed air to generate combustion gas. As shown in the examples, two can co-burners (can co
mbustor) 60 is an air trap through the wall of a double layer turbine inlet manifold 62 (only a single combustor is shown in Figure 1).
Connected with 54. Preferably, during steady state operation, all of the compressed air exiting the second compression stage 40 is passed to the combustor 60 via the air pocket 54 and the walls of the manifold 62. It may be used for partial cooling or sealing purposes.

Still further in accordance with the invention, the gas generator includes a single stage radiant inlet turbine having an inlet and an outlet. The turbine is connected for direct drive to the shaft assembly to which the compression stage is mounted and is also connected to the combustion means for partial expansion. As shown in the example, the gas generator 10 includes a blade 70 having a blade tip 70a.
A radiant inflow turbine 66 having a turbine rotor 68 having. Turbine 66 passes from combustor 60 through turbine inlet nozzle assembly 77 and turbine inlet manifold 62.
To receive the high temperature combustion gas. Turbine rotor 68
Is directly connected to the shaft assembly 20 for rotating both the rotor 58 of the high pressure compression stage 40 and the roller holder 18 of the low pressure compression stage 14.

Further in accordance with the present invention, the gas generator includes exhaust means connected to the turbine outlet to direct the partially expanded combustion gas for further work expansion.
As shown in the exemplary embodiment, the gas generator 10 receives the partially expanded combustion gas from the turbine 66 and may, for example, be a downstream free power turbine unit.
nit) 90 or a manifold 74 connected to feed to a free jet thruster 92 (FIG. 1A). The free power turbine unit 90 can have one or more individual stages (two stages are shown in FIG. 1).

FIG. 2 shows a gas generator 2 according to the present invention.
Specific velocity curves for the two compression stages and the radiative turbine stage are shown. A gas generator has an inlet Mach number of 1.4 or more.
Using a dual introduction centrifugal compressor with a 9: 1 pressure ratio and a second stage single introduction centrifugal compressor with a pressure ratio of about 3.5: 1 (ie 50% or less of the pressure ratio of the first stage), a 20: 1 It has the above total pressure ratio. As can be seen in the figure, the efficiency of the radiant-flow turbine peaks near the efficiency peak of the second compression stage when they all rotate together coaxially as part of the gas generator. This means that at a compressor specific speed of about 0.65 to 0.85 and a turbine specific speed of about 0.50 to 0.75 obtained at a radial inflow turbine specific speed of about 0.50 to 0.75.
It is represented by the area shown by A in FIG.

Region B is a second stage centrifugal compressor followed by one or more axial flow stages or a single inlet centrifugal compressor as a hypothetical alternative low pressure first stage made by the teachings of prior art high pressure ratio gas generators. It is shown that the optimum region causes a significant drop. FIG. 2 also shows that when the radiant inflow gas generator turbine is to be matched to the low specific speed compressor section, it falls below the optimum value. Therefore, all the rotation components of the gas generator flow passage according to the present invention rotate at a high optimum specific speed, the physical size of the rotor is reduced, and the manufacturing cost is reduced. Further, the inertia of the rotor assembly is reduced, which is convenient for quick start.

The double introduction first stage compressor has twice the mass flow as the single introduction compressor of comparable size. Under steady-state operating conditions, when air is compressed to design criteria in the first stage,
The latter flow path shape is the total first stage compressed gas flow rate
Harmonize with. However, during start-up and before the first stage, the volumetric flow rate is reduced to the design standard and the second stage compressor cannot accommodate a volumetric flow rate greater than the design value.

Thus, means are preferably formed for outflowing the compressed air exiting the first compression stage during start-up conditions, the outflowing air avoiding the second compression stage. As shown in the embodiment of FIG. 1, the outflow duct 76 has a first compression stage when the valve 78 is activated during startup.
It is connected to a cross duct 44 to receive compressed air from 14. An automatic controller 80 controls the valve 78, but manual operation can be switched instead of the controller 80. Those of ordinary skill in the art will appreciate the controller 80 shown herein.
It is possible to easily select an appropriate device for forming the function of. The effluent air can simply be vented to the atmosphere as shown in FIG.

The drawings of the present invention and the prior art are sufficiently different. Absolute range of pressure ratio split between first stage and second stage compressor coupled with specific speed, total pressure ratio and mechanically and aerodynamically matched drive elements, ie radiant gas generator turbine Regarding (absolute range), it is possible to obtain a compression ratio of 20: 1 or more with a turbine engine of 1000 or less, for example. Meanwhile, "Aeronautical Journal of the Royal Aeronautical S (Aeronautical Journal of the Royal Aeronautical S
January / February 1984 "Rolls Royce European Symposium" entitled "Small Engine Technology" Philip G Ruffle
s), the most expected pressure ratio for engine size has hitherto been 14: 1 to 16: 1, as evidenced by his paper.

For example, the stagnant temperature of the combustion gases impinging on the turbine blade tip 70a above about 700 m / s of the present invention, and therefore the turbine metal temperature, is less than the axial turbine stage temperature exposed to the same nozzle inlet temperature. Much lower (less than about 100-200 ℃). Thus, the high blade tip velocities of the inventive fluent turbine, dictated by the compressor requirements, actually form an additional advantage in that the metal temperature is lower. This allows for high turbine inlet temperatures which, together with the high pressure ratio of the present invention, results in very low fuel consumption rates.
The high pressure ratio cycle of the present invention is about 12% current material without introducing cooling to the turbine rotor.
High combustion efficiency can be achieved at combustion temperatures up to 00 ° C. At this temperature level, all existing axial turbines require cooling. Cooling the radiant turbine when the gas generator according to the invention uses a pressure ratio of 20: 1 or higher,
Alternatively, it is preferable to use a non-metallic material for the rotor, which can further improve efficiency. Cooling also increases the specific efficacy even if the thermal efficiency is not improved.

As is currently contemplated, the gas generator shown in FIG. 1 theoretically operates at a total pressure ratio of about 20: 1 and has an air flow of about 4.0 lb / s and a turbine inlet temperature of about 2200 ° F. and above. Produces 700 equivalent shaft power, has a thermal efficiency of about 35%, about 5.0:
Use a gas generator turbine expansion ratio of 1. The turbine rotor speed is about 92000 rpm, which is equivalent to the turbine blade tip speed of about 750 m / s. The designed equivalent engine fuel consumption will be about 0.35-0.40 lb // h. The specific weight of the engine using this gas generator is only about 10% of that of the comparable diesel engine.

Conventionally, the thermal efficiencies cited above are very large (eg, over 30,000) gas turbine engines using axial flow components or recuperators / regenerators with low pressure ratios.
Can only be obtained with small power units using perator / regenarator). Of course, conventional diesel engines show good thermal efficiency, but they show considerable disadvantages in terms of size and weight.

It will be apparent to those skilled in the art that various modifications and variations can be made in the gas turbine gas generator system of the present invention without departing from the scope and spirit of the invention.

Claims (16)

[Claims] 1. (a) (i) a first-stage double-introduction centrifugal air compressor having a pair of inlets and one common outlet, (ii) a flow-connected one with the first-stage common outlet. A second stage centrifugal air compressor having two inlets and a second stage outlet and adjacent to the first stage compressor; and (iii) the first stage and the second stage for rotating at the same angle. A compressor device that forms a total pressure ratio of greater than about 15: 1, including a shaft assembly that mechanically connects the two components; (b) compressed air and combustion fuel that uses compressed air to generate combustion gases. Combustion means operatively connected to said second stage outlet for receiving; (c) a one stage radiant inflow turbine having one inlet and one outlet, said turbine operatively connected to said shaft assembly drive and combustion Flow coupled to the combustion means to partially expand the gas (D) in a dual-introduction radial turbine gas generator including exhaust means connected to the turbine outlet for supplying partially expanded combustion gas, (i) the pressure ratio of the first compression stage is Is greater than twice the pressure ratio of the second compression stage, and (ii) the comparative rotational speeds of the first compression stage and the second compression stage are close to their optimum values and are greater than about 0.60, respectively. A dual-introduction radial turbine gas generator characterized by: 2. The gas generator according to claim 1, wherein the specific speed of each of the first compression stage and the second compression stage is between about 0.65 and about 0.85. 3. A gas generator according to claim 1, wherein the total pressure ratio of the first compression stage and the second compression stage is greater than about 20: 1. 4. A gas generator according to claim 1, wherein the first compression stage pressure ratio is between about 6: 1 and 9: 1. 5. The gas generator of claim 1 wherein the second compression stage pressure ratio is between about 2: 1 and 4: 1. 6. A gas generator according to claim 1, characterized in that almost all the compressed air leaving said second compression stage is received by said combustion means. 7. A first compression stage and a second compression stage are coaxial with said radiant inflow turbine, said turbine operating at a peripheral speed greater than about 700 m / s. The gas generator according to item 1. 8. The Mach number at the tip of the inducer is about 1: 4 or more with respect to the Mach number at the inlet of the first stage double introduction compressor, as set forth in claim 1. Gas generator. 9. A gas generator according to claim 1, further comprising means for discharging a portion of the compressed air leaving said first compression stage during start-up conditions. 10. The second aspect of the present invention, wherein the specific speed of the radiant flow turbine is between about 0.50 and about 0.75.
The gas generator according to the item. 11. A gas turbine engine having a gas generator according to claim 1 in combination with a free power turbine. 12. The first ratio of claim 1, wherein the specific speed of the radiant inlet turbine is between about 0.50 and about 0.75.
The gas generator according to the item. 13. The specific speed of the first and second compression stages is 0.65.
2. The gas generator according to claim 1, wherein the specific speed of the radiant-flow turbine is between 0.50 and 0.75. 14. The gas generator according to claim 13, wherein the total pressure ratio of the first compression stage and the second compression stage is greater than about 20: 1. 15. The gas generator of claim 13 wherein the first compression stage pressure ratio is between about 6: 1 and 9: 1. 16. The gas generator of claim 13 wherein the second compression stage pressure ratio is between about 2: 1 and 4: 1.