JPH03290024A - Two-shaft type intermediate pressure power turbine regeneration type gas turbine - Google Patents

Abstract

PURPOSE:To improve part load thermal efficiency of a two-shaft regeneration type gas turbine by changing the turbine arrangement to maintain the energy distribution ratio of a compressor driving turbines and a power turbine constant regardless of load factor, and automatically controlling the flow rates of the turbines interrelatedly. CONSTITUTION:Compressor driving turbines for driving a compressor C are composed in two stages of a high pressure turbine HT and a low pressure turbine LT to employ an intermediate pressure turbine IT as a power turbine, and the IT is not mechanically coupled to the gas generator shaft. The intermedi ate pressure (power) turbine IT is rotatably arranged on the gas generator shaft to take out the work to the outside by means of a gear train G. The power turbine shaft is provided separate from the gas generator shaft, and a compressor driving low pressure turbine LT is arranged in the downstream of the power turbine IT. The LT is not mechanically coupled to the power turbine shaft to be coupled to the gas generator shaft by means of the gear train G. Thus, the high pressure turbine HT and the low pressure turbine LT have the same rotational speed and a constant rotation ratio.

Description

[Detailed Description of the Invention] [Field of Industrial Application] The present invention relates to a two-shaft regenerative gas turbine suitable for use in vehicles, ships, and mechanical drives. First, it can be improved by simply changing the turbine arrangement.

[Conventional technology]

In a two-shaft gas turbine, where the output turbine can rotate freely relative to the gas generator, the n-positive torque is approximately twice the rated point, so
Suitable for vehicles, ships, and mechanical drive engines that require high torque at low speeds.

However, in a two-shaft gas turbine, the energy distribution between the compressor drive turbine and the output turbine changes depending on the load factor, and the energy distribution to the output turbine decreases at low loads, so even if a regenerative heat exchanger is used, Part load thermal efficiency is poor.

In order to improve the part-load thermal efficiency of the twin-shaft N-type gas turbine, a variable 1111 ellipse is used as a means of adjusting the energy distribution between the high and low pressure turbines.

There are three types of conventional variable mechanisms:

) Variable output turbine In blade The nx at the output turbine inlet is made a variable blade and is throttled at 1 partial load to reduce the output turbine flow rate and prevent the output turbine expansion ratio from decreasing during machining.Therefore, the control method is This is the flow rate control of the output turbine.

ii) Insert the power transmission mechanism between the power transfer machine lateral output shaft and the gas generator shaft,
The maximum cycle temperature is increased by using a single shaft under partial load. Therefore, the control method is to control the torque transmitted between the shafts.

iii) KTT 3m gas turbine The KTT (Kronogard Turbine) is being developed by the Swedish company United Turbine.
Bine Trans+*1sson＞3-shaft gas turbine, 2-shaft gas turbine. An auxiliary turbine is installed after the output turbine, and the output shaft and compressor shaft are driven through planetary gears. This configuration gives the engine a transmission function. Achieves low fuel consumption and high torque ratio under load. This method requires variable vanes for matching during partial loads. Therefore, the control method is to control the torque transmitted to the output shaft and the compressor shaft, and to control the output turbine flow rate.

[Problem to be solved by the invention]

Conventional OK'Rt! The disadvantages of the control method using one ellipse are as follows.

) Control of output turbine flow rate If the pressure ratio is increased by throttling the outlet side at the rotation speed (on the head curve) of the compressor, the compressor will approach the surging region. Therefore, the output turbine flow rate control method that throttles the compressor outlet side using variable stator vanes cannot significantly improve fuel efficiency due to the surge limit of the compressor.

Furthermore, providing a variable mechanism in a high temperature and high pressure section is not a desirable method because it reduces the lubricating power, causes air leaks, and increases complexity.

ii) Control of Transmitted Torque Control of the transmitted torque between the shafts requires a slipping clutch, or a multi-disc clutch consisting of a planetary gear and a transmission element, etc. to be provided in the high-speed rotating section, and this also has a complicated structure.

The present invention has been made in view of the above points.

The problems to be solved by the present invention are listed below.

) The object of the present invention is to significantly improve the partial load maturation efficiency of a two-shaft regenerative gas turbine, and at the same time to simplify the improvement means.

11> Compression 8! to improve partial load silent efficiency! The energy distribution ratio between the drive turbine and the output turbine is made constant regardless of the load factor.

m) Eliminate complicated variable mechanisms to simplify improvement measures;
By simply changing the turbine arrangement, the flow rates between the turbines are automatically controlled, and the energy distribution ratio is kept constant.

[Means to solve problems]

The structure and cycle of the present invention are shown in FIG. 1121 and FIG.

In the figure, C is pressure lll5. HT stands for high pressure turbine, IT
is an intermediate pressure turbine, LT is a low pressure turbine, CC is a combustor, HE is a regeneration mature exchanger, and G is a gear train P is a load.

In the present invention, the compressor driving turbine is made up of two stages: a high pressure turbine HT and a low pressure turbine LT. The output turbine is an intermediate pressure turbine IT, which is not mechanically connected to the gas generator shaft. In the configuration shown in FIG. 1, the intermediate pressure (output) turbine IT is rotatably placed on the gas generator shaft, In the configuration of FIG. 2, in which work is extracted externally by row G, the output turbine shaft is provided separately from the gas generator shaft, and a compressor-driven low-pressure turbine LT is placed downstream of the output turbine IT.

LT is not mechanically coupled to the output turbine shaft, but is coupled to the gas generator shaft by a gear train G. Therefore, the high pressure turbine HT and the low pressure turbine LT have the same rotation speed or constant rotation ratio. Therefore, in the case of Figure 2, there are not 3 axes, but
It is essentially two axes.

FIG. 2 and the KTTB shaft gas turbine are shown in FIG.
There is no star gear system, the low pressure turbine LT is connected only to the gas generator shaft, and the variable stator blades are outstanding. Therefore, the control method is different, and in the present invention, the variable mechanism is eliminated and mutual flow rate control of the turbines is automatically performed.

To put it simply, the present invention is a gas turbine that uses a regenerative one-wheel three-stage turbine medium-pressure turbine as a free turbine and has two shafts, from which output is extracted.

[Effect]

The operating principle of the present invention is that by connecting the shafts of the high-pressure turbine HT and the low-pressure turbine LT, the high-pressure The goal is to keep the energy distribution ratio to the three turbines, medium pressure and low pressure, constant.

If the three turbines of the present invention are made to have similar structures, the law of similarity will hold that the performance of the turbines is that the flow rate is proportional to the rotational speed, and the adiabatic head is proportional to the flow rate.

Since the high-pressure turbine IT (which is a compressor-driving turbine) and the low-pressure turbine LT are mechanically coupled, the rotational speed or rotational ratio is constant regardless of load fluctuations. Since the sizes of both turbines are fixed, the flow rates of both the high and low pressure turbines are equal even under partial load, and the ratio of the adiabatic heads is also constant regardless of the load factor due to the law of similarity. That is, if the ratio of high-pressure turbine work W□ and low-pressure turbine work WL is α, then α=WL /Wll =const
If the flow rates of both the high and low pressure turbines are equal ( ), then the flow rate of the intermediate pressure (output) turbine IT sandwiched between them is also equal to this, and if the flow rates are equal, the internal theoretical calculation head (energy ) also has a constant distribution ratio regardless of the load factor. Therefore, the energy distribution ratio to the three turbines is also constant regardless of the load factor. Therefore, if the turbine adiabatic efficiency is constant, the following equation holds true, where β is the ratio of the high-pressure turbine work W and the intermediate-pressure (output) turbine work WI.

β=W, /W, -8°. st
(2) With the above as the basic principle, the operation of the present invention will be explained based on theoretical calculations. The symbols and setting values used for calculation are shown below.

Wo Compressor work W ca Compressor drive turbine work W, High pressure turbine work W, Medium pressure (output) turbine work W, Low pressure turbine work Wo Specific output Q, Combustor supply heat amount η Ripe efficiency Lo Load factor rc Pressure ratio r? Turbine total j1 tension ratio r, High pressure turbine 1M tension ratio r, Medium pressure (output) turbine expansion ratio rL, Low pressure turbine expansion ratio α = WL/W, Work ratio of low pressure and high pressure turbine β-
W, /W, Ratio of work between intermediate pressure and high pressure turbine G Pressure i Machine flow rate G, l High pressure turbine flow rate GL Low pressure turbine flow rate S, l High pressure turbine nozzle area SL Low pressure turbine nozzle area Specific heat ratio of air To c-14 combustion The specific heat ratio of the gas is t = 1. :'l:'1
mc= (to c-1>/to C m t = (to t-1)/to constant pressure specific heat of the compressor Cpc= 0.240kca1
% kg Turbine constant pressure ratio % Cpt = 0
, 276 k c a l 7 k g gas constant
R = 29.27kgm/kBK regenerative silent exchanger air side pressure loss ΔP A = 0.02P.

Heat exchanger combustion gas side pressure loss △P c= 0.02P
.

Combustor pressure loss Δp ee = 0.03
F1.

Compression II shearing efficiency ηc = 0.82 Turbine adiabatic efficiency η, = 0.85 Combustion efficiency
ηcc = 0.99 Regenerative exchanger temperature efficiency η,, -0.85 Mechanical efficiency
η, = 0.97 To simplify the calculation, we omit the change in the flow rate of the working fluid due to the addition of fuel, and assume that the specific heat and specific heat ratio are constant in the compressor and turbine sections, and the adiabatic efficiency of the compressor and turbine is also Each is constant. The cycle calculation formula of the present invention is shown below.

Wc=CpcT1(r”-1)/ηc(3)W11=
77? CF, T, (] -,, -at)
(4) w+=ηtrp, t, (1-1
−1−”) (5) W and ηt
Cr+Ti (1-rL-”)
(6)W
? =η? CPiT+(1,,-m')(7)α-w,,
'w, from wc, ww+wL= (1+α)W++
(8) WO-η・W・
(9) Q+=CPc (
T*-T3) /ηee (l
o)7. -W. 'Q, (II) Pi=P(re Ps=P2-ΔPII P,=Pj-ΔP,c=P,-(ΔPA+ΔP-c)P
s=P,/r*PG=Ps/r+P,=Pg/rL=Pm+ΔP.

r' s = P (12><13) (10 (15) (16> (17) <18> T, -T, il+ (r″″'-1)/77-1T3=
(1-η,)T2+η,,T.

Ts-Tt (1,-77741-r++-”) IT
s=Ti 11-77? (1-r,-++t) )T
I-TI (177T (1rt-1)IT・=(
1-η5t) T++ ηMET2 * Calculation of design point performance The following assumptions are made at the design point.

1.033kB in P, /cm' rc=5T, =2
88K T, = 1200KT TS, T +
-r c, high pressure turbine 1 tension (20) (2]) (22) (23) (24) (25) The ratio r8 can be found from η, Wc, and c=We. In the formula on the left (
8) Substituting the formula, Wc=η, (1+α)W, l Therefore, CpcTI(r” 1 ) = 77m'7c'7yCpt'L(1-rx-") (
1-+α) (26) Next, medium pressure (output)
Determine the turbine expansion ratio r+.

From WT=Wcc+W+, W+=Wt-Waa=Wy-(1+a) WN77
? CPI75 (1-rr-”) = ηyCpt'
L<1-rt--t)77yCptT+
(1r ut) (1+α) (1-η clove η?r□-”
) (1r, Su゛〉-(1-r?Su゛) (1r
u-”)(1+α,)low-pressure turbine expansion ratio r, is (19
) From the formula, rL−rD/ (rHrI)

(29) Using the above equation, α−0,
8, the F tension ratio of each turbine at the design point is r++#1.53 r+! ;] 925 rL
”;1.583 turbine total I!ij tension ratio is rt#4.6
It becomes 6.

If you know the Wr3 tension ratio of each turbine, (20) to (25)
The temperature in each state can be determined from the equations, and the thermal efficiency η and specific output W0 at the design point can be obtained from equations (3) to (11).

The calculation result is η#o, 348 We#37.463keml/
kg. Also, if β is determined at the design point, β=W
, /W, # 1, 37 r at the design point of the conventional two-axis regeneration cycle calculated under the same calculation conditions. , rr and η, wo are r11?2.2
27 rL=2.092η#0.36 woL,
3s, 698 kcal/kg In the present invention, since the compressor driving turbine has two stages, IT and LT, the compressor driving turbine expansion ratio (r MX r L
) is larger than the conventional two-shaft regeneration cycle, and the output turbine expansion ratio r is correspondingly smaller, so η and wp at the design point are slightly lower than in the conventional two-shaft regeneration cycle.

Calculation of Partial Load Performance Next, we will show the matching calculation at partial load.First, we will calculate the tension ratio of each turbine 11 assuming the pressure ratio re (gas generator rotational speed) at partial load.

The work of the total turbine is the sum of the work of the high, medium, and low pressure turbines, so 5 W, -W. +W, +WL (30) (30> Formula is a formula that indicates work per unit flow rate, so
This means that the three turbine flow rates are matched. Substituting equations (1) and (2) into equation (30), W, = (1+α+β)W,
(3]) Substituting equations (4) and (7) into equation (31), rzycr+T+ (]-rtp) =
ηTCPITI (1+a+β) (1-r, -”) The intermediate pressure (output) turbine expansion ratio rI is W in equation (2), = βW
H riff 7 y Cp + dingh (J rrsl) = β77y
cr+T1 (] rr*”') The low pressure turbine expansion ratio rL is obtained from equation (29). The cycle maximum temperature T for each pressure ratio at the primary partial load is the output matching equation, and at the same time, Equation (26), which means that the flow rates of the compressor and the compressor-driving turbine are equal, is transformed into (32).
) is obtained from the equation (34), and T4 for each pressure ratio is determined from the equation (34), the maximum cycle temperature that simultaneously satisfies the matching conditions of flow rate and output can be obtained.

If r, r'+, rL, and T for each pressure ratio are known, then the thermal efficiency η is 2 from the equation (11) as described above.The specific output W0 can be obtained from the equation (9).

Next, the compressor flow rate G for each pressure ratio at partial load is determined from the turbine flow rate. To simplify the calculations, it is assumed that the recoil of each turbine is zero and that full stage expansion occurs at the nozzle. If the critical pressure ratio of combustion gas is 0542, then at the high-pressure turbine outlet temperature, when (1/r,)<0.542, when (1/r,)>0.542, even on the low-pressure turbine LT side ( 1/rL)<0.542 When (1/r,)>0.542, calculate the high pressure turbine nozzle flow rate G8 and the low pressure turbine nozzle flow rate GL. Therefore, at the design point, the unit weight flow rate By calculating the respective turbine nozzle areas S□, S, it is possible to calculate the high-pressure turbidity a Mc- and the low-pressure turbine flow iGL for each pressure ratio.

When calculating G,,GL, at partial load, G□<
Becomes GL. Since the high-pressure turbine nozzle inlet state is determined by the matching conditions, more gas than the flow rate calculated by equation (35) cannot pass through the high-pressure turbine nozzle. Therefore, the compression W1 flow rate was set as G-GII.

The load factor L0 for each pressure ratio is determined from the following equation.

1-o=w0 η and Lo shown by dotted lines in Figure 3 are for the conventional two-shaft regeneration cycle obtained under the same calculation conditions.From Figure 0, it is clear that the present invention significantly improves the partial load thermal efficiency. I understand.

Further, FIG. 4 shows the turbine inlet temperature T and outlet temperature T of the present invention as solid lines, and the turbine inlet temperature T and outlet temperature T6 of the conventional two-shaft regeneration cycle as dotted lines.
It can be seen that the difference in thermal efficiency η in the figure is caused by the difference in the maximum cycle temperature T.

However, in the present invention, when the pressure ratio becomes r c < 2, the turbine outlet temperature T7 increases, which may cause a problem with the material of the regeneration ripening exchanger HE.
Maybe I should keep the temperature down.

Next, FIG. 5 shows the high pressure turbine flow rate ratio G for each pressure ratio in the present invention (that is, the compression m flow rate ratio G) and the low pressure turbine flow rate ratio GL, which clearly shows that G ++ < G L
It is.

However, in reality, the high pressure turbine flow rate and the low pressure turbine flow rate must be equal due to the principle of continuity. Each pressure ratio @nr derived to understand the partial load performance of the present invention. rl,
r, and T simultaneously satisfy the matching conditions for flow rate and output. Nevertheless, G s <-, G L
becomes.

This turbine flow rate calculation shows that when the compressor drive turbine is made into two stages, the axial flow velocity (flow coefficient) decreases on the low pressure turbine side at partial load due to the aerodynamic similarity law of the turbine. The flow rate decreases from GL to G□.

This flow phenomenon at partial load of the turbine is
A multi-stage compressor has a small flow coefficient near the first stage at low speed.
This is exactly the opposite of what happens when a positive stall occurs.

From the above, the present invention has a two-stage compressor-driving turbine with an output turbine in between, and can automatically control the gas flow within the turbine stage without using a variable mechanism. Matching occurs where the output turbine expansion ratio rI is high.

In other words, the matching conditions of the present invention under partial load are similar to those of a single-shaft gas turbine, rather than those of a two-shaft gas turbine.

η and η shown in FIG. The calculation is performed by changing the pressure ratio me while keeping the turbine adiabatic efficiency η1 and mechanical efficiency η constant. Therefore, in FIG. 3, the peak efficiency and peak output at each rotation speed of the gas generator of the present invention and the conventional two-shaft regeneration cycle are compared. In other words, FIG. 3 corresponds to a comparison of the lowest fuel efficiency rates for each load factor.

When the gas generator is at the rotational speed of Since the efficiency of the turbine decreases, the thermal efficiency and power output of the entire combo unit decreases.

In order to calculate the performance by changing the output turbine rotation speed while the gas generator rotation speed is constant, it is necessary to use the characteristic curve of a specific compressor and a specific turbine, so we will omit it here.

〔Effect of the invention〕

Taking the thermal efficiency near 20% load factor from Figure 3 as an example, in the conventional two-axis regeneration cycle, when L0 = 0.196, η = 0.2047, and in the present invention, when L0 = 0.197, η =0.2773 Therefore, the difference in thermal efficiency is 0.0726 Therefore, the present invention can improve the partial load thermal efficiency of a two-shaft regeneration cycle by about 35.5 p; at a load factor of around 20%.

In a conventional fixed Wpr & two-shaft gas turbine, if it is attempted to maintain a constant energy distribution ratio to the high and low pressure turbines, the maximum cycle temperature will be extremely excessive under partial load, making it impossible to achieve this. Furthermore, if this were to be achieved using the conventional variable output turbine stationary blades, the output turbine flow rate would have to be reduced by a large amount, as seen from FIG. 5, which would go against the compressor characteristics.

From the above, the present invention makes use of the similarity law of turbines to make the flow of three turbines similar to that of a single-shaft gas turbine, thereby making it possible to keep the energy distribution ratio to each turbine constant. The partial load thermal efficiency of a two-shaft regenerative gas turbine can be improved to a greater extent than ever before without using a lateral shaft.

[Brief explanation of drawings]

Figures 1 and 2 are diagrams showing the configuration and cycle of the present invention, and Figure 3 shows the pressure ratio r of the present invention and the conventional two-shaft regeneration cycle.
Force efficiency η and load factor ri for e. Figure 4 shows the relationship between the turbine inlet temperature T and outlet temperature (T6, T,) of the present invention and the conventional two-shaft regeneration cycle.
Figure 5 shows the nozzle/L flow rate. A diagram showing the calculated high pressure turbine flow rate ratio G8 and low pressure turbine flow rate ratio GL. In Fig. 1 and Fig. 2, C compressor, HT: high pressure turbine, IT: intermediate pressure (output) turbine, LT: low pressure turbine, CC: combustor, l: IE: regenerative heat exchanger,
G gear train, P load.

Claims (1)

[Claims] A high-pressure turbine and a low-pressure turbine in turbine stages consisting of high-pressure, intermediate-pressure, and low-pressure turbines are mechanically coupled to a compressor to form a compressor-driven turbine, and the intermediate-pressure turbine is rotatable about the gas generator shaft to form an output turbine, which compresses A two-shaft type that connects the machine outlet and compressor-driven high-pressure turbine inlet to a regenerative heat exchanger and then a combustor, and sends the gas exiting the compressor-driven low-pressure turbine to the regenerative heat exchanger to recover exhaust heat. Medium pressure power turbine regenerative gas turbine.