JPS6250663B2 - - Google Patents

Description

[Detailed description of the invention] The present invention relates to a single speed reversible two-stage pump water turbine,
In particular, the present invention relates to a single-speed reversible two-stage pump water turbine that can perform high-performance operation during steady operation.

In recent years, there has been a trend to construct pumped storage power plants at locations with high heads, and the demand for multi-stage pump turbines is increasing. There is an increasing demand for single-speed reversible two-stage pump water turbines, which are structurally relatively easy to provide with a mechanism.

However, in a conventional two-stage pump-turbine in which the runners of each stage are constructed from runners with the same outer diameter and hydraulically similar, there are problems with the hydraulic characteristics peculiar to single-speed reversible pump-turbines, which will be described later. Therefore, there was a problem in that either the water turbine operation or the pump operation had to be performed at each step in a region of low hydraulic performance far from the highest efficiency state.

SUMMARY OF THE INVENTION Therefore, an object of the present invention is to provide a single-speed reversible two-stage pump-turbine that can improve the overall hydraulic performance of each stage during steady operation.

According to the present invention, the above object is achieved by connecting the high-pressure stage section and the low-pressure stage section in series through a return passage, and connecting the runners of each stage section to a single main shaft of the water turbine. In a single-speed reversible two-stage pump water turbine that operates, a movable guide vane that can change the opening of the water port is provided on the outer periphery of the runner in each stage, and the outer diameter of the runner blade of the runner in the low-pressure stage is the same as that in the high-pressure stage. This is achieved by setting the diameter within the range of 0.86 to 0.96 times the outer diameter of the runner blade.

Embodiments of a single-speed reversible two-stage pump turbine according to the present invention will be described below with reference to the drawings.

Figure 1 shows the structure of a single-speed reversible two-stage pump turbine, in which a high-pressure stage runner 2 and a low-pressure stage runner 3 are fixed on the shaft of a single main shaft 1 at a distance in the axial direction. . The above high pressure stage runner 2 has an upper cover 4
and a lower cover 5, while the low pressure stage runner 3 is surrounded by an upper cover 6 and a lower cover 7,
A high-pressure stage runner chamber 8 and a low-pressure stage runner chamber 9 are configured. The high pressure stage runner chamber 8 and the low pressure stage runner chamber 9 are connected through a return passage 10.

Further, a spiral casing 11 is disposed outside the high pressure stage runner chamber 8, and the spiral casing 11 is in communication with the high pressure stage runner chamber 8.
The inlet of is connected to a penstock 14 via an inlet valve 13.

Furthermore, a movable guide vane 15 is provided on the outer periphery of the high-pressure stage runner 2, and a movable guide vane 15 that can change the water port opening is provided on the outer periphery of the low-pressure stage runner 3. It is arranged in a shape. These movable guide vanes 15 and 16 are connected to an operating mechanism (not shown) and are operated steadily by a control device.

When operating the single-speed reversible two-stage pump turbine configured as described above, water flows from the penstock 14 into the spiral casing 11 connected to the penstock 14 with the inlet valve 13 open, and this water flow are the movable guide vane 15 of the high-pressure stage section and the high-pressure stage runner 2.
It passes through the return passage 10 and further flows through the movable guide vane 16 of the low pressure stage section and the low pressure stage runner 3,
The water flows through a suction pipe 17 connected to a water discharge channel (not shown). On the other hand, in the case of pump operation in which the runner rotates in the opposite direction at the same rotational speed as the water turbine, the water flow pumped by the low-pressure stage runner 3 is sucked out through the reverse route as in the case of water turbine operation. tube 17
The water then flows to the penstock 14.

In this way, the overall hydraulic characteristics of a single-speed, reversible, two-stage pump-turbine in which each stage is connected in series through a return passage can be determined by combining the hydraulic characteristics of a single-speed, reversible, single-stage pump-turbine in each stage. Given. Therefore, when considering the problems with the hydraulic characteristics of a single-speed reversible two-stage pump-turbine, it is first necessary to accurately understand the problems with the hydraulic characteristics of a single-speed reversible single-stage pump-turbine. Therefore, we will consider it next.

Generally, in a single-speed reversible single-stage pump-turbine, the rotational speed of the pump-turbine is N (rpm), the effective head of the turbine is Ht (m), the efficiency ηt, the total head of the pump is Hp (m), and the efficiency is ηp. , the envelope of the efficiency curve of each guide vane opening with respect to unit rotational speed N/√, N/√ for a water turbine and a pump is shown in FIG. 2 as an example. That is,
As shown in Figure 2, the unit rotational speeds N/√ ₀ and N/√ ₀ , which give the highest efficiency points for water turbines and pumps, do not match, and N/√ ₀ is always larger. This is one of the problems in hydraulic characteristics that cannot be avoided in pump turbines. Efforts have been made to make this difference as close as possible, but in practice it usually falls within the following range.

In other words, as shown in Figure 2 and Equation (1) above, in a single-speed reversible pump-turbine, the maximum efficiency point of the pump is far away from the maximum efficiency point of the turbine in terms of unit rotational speed ratio. , the efficiency during pump operation at a unit rotational speed corresponding to the maximum efficiency point of the water wheel is extremely low, and similarly, the efficiency of the waterwheel at a unit rotational speed corresponding to the maximum efficiency point of the pump is also extremely low. For this reason,
At the standard operating water level of a pumped storage power plant, if the operating conditions of the pump-turbine are determined so that the pump operates at or near its maximum efficiency, as is normally done, the condition shown in Figure 2 and equation (1) is Due to the constraints related to hydraulic characteristics, the water turbine is operated under _a high unit rotational speed state far from the maximum efficiency. In other words, the operating head is equivalent to the maximum efficiency state. You will have no choice but to choose. In this way, when considering the hydraulic characteristics of a pump-turbine, the unit rotational speed (N/√,
N/√) has an important meaning, and next we will discuss the correlation between this unit rotational speed and the outer diameter D of the runner blade, which basically controls the hydraulic characteristics of the pump-turbine.

Generally, at the highest efficiency point of the pump, which represents the design point of the hydraulic characteristics of a pump-turbine, the rotational speed of the runner is N (rpm), the outer diameter of the blade is D (m), the outer circumferential speed of the blade is u (m/s), The blade peripheral velocity coefficient is φ, the pump total head is Hp ₀ (m), and the gravitational acceleration is g (m/
s ² ), the following equations (2) and (3) are given hydraulically.

u=πDN/60=φ√2 ₀ …(2) or φ=(π/60√2)(N/√ ₀ )D = (1/√2)(u/√ ₀ …(3) On the other hand , the flow rate at the pump's highest efficiency point is Qp (m ⁸ /
s), the specific speed Ns, which is the basic condition for pump-turbine design, is Ns=N・(Qp) ^1/2 / (Hp ₀ ) ^1/3
given as. According to the results of a model hydraulic characteristic test conducted by sequentially and systematically processing the outer diameter D of the runner blades in a pump turbine designed under the condition of a certain specific speed Ns, the outer diameter D of the blades
The unit rotational speed N/√ ₀ increases as the unit rotational speed N/√ 0 decreases, and within the range where practical performance can be obtained, the following (4)
As shown in the formula, it has been confirmed that the two are in a reciprocal proportional relationship.

(N/√ ₀ )・D=constant ...(4) This means that if the blade outer diameter D is made small in a pump turbine designed under the condition of a given specific speed Ns, it will be within the range where practical performance can be obtained. Therefore, it can be said that the blade outer circumferential speed coefficient φ in the above equation (3) hardly changes, and therefore the result brings about the relationship in the above equation (4).

In addition, when the outer diameter D of the blade is reduced in this way, the unit rotational speed N/√ of the pump maximum efficiency point is
₀ , the unit rotation speed N/√ ₀ at the highest efficiency point of the water turbine also increases, and each unit rotation speed ratio is within the range defined by equation (1) above, and the hydraulic characteristics of both the pump and the water turbine are It is possible to shift the overall rotational speed to the high unit rotational speed side, that is, to the low head side.

Next, based on the above-mentioned results of the discussion on the basic problems in the hydraulic characteristics peculiar to single-speed reversible single-stage pump turbines, we will refer to Fig. In the case of D ₁ = D ₂ in Figure 1), we will discuss the problems with the hydraulic characteristics during steady operation in a normal single-speed reversible two-stage pump-turbine constructed from hydraulically similar units. In this case, the relationship between the hydraulic characteristics of each stage during steady operation of the water turbine is shown in FIG. 3. In Figure 3, H ₁ is the effective head of the low-pressure stage, H ₂ is the effective head of the low-pressure stage, Q is the flow rate, and those with an index 0 are at the reference operating state point (□0 in the figure). value, a ₀ is the opening of each stage water port (guide vane opening) during the standard operating state, and a with a positive integer subscripted to a (for example, a ₁ ) is the opening when the opening is more than the standard opening a ₀ . The opening of the stepped water port, a with a negative integer subscripted (for example, a _-1 ), is the opening of each stepped water port when the water opening is slightly opened according to the standard opening, and △η is the relative difference from the maximum efficiency of the turbine. This is the relative efficiency difference between the two turbines. Figure 3 shows the flow rate ratio Q/Q ₀ on the horizontal axis.
In addition, the vertical axis shows the effective head ratio of the high pressure step and low pressure step.
H ₁ to H ₁₀ and H ₂ /H ₂₀ are used to display the relationship between the hydraulic characteristics of the effective head and the flow rate for each step. Therefore, the overall effective head as a two-stage pump turbine is It is given by adding up the effective heads of each step. In FIG. 3, the inlet valve 13
When the water port opening of the movable guide vanes in each stage is controlled simultaneously during steady operation by fully opening the runners, as mentioned above, the runners in each stage should be configured with vanes that have the same outer diameter and are hydraulically similar. Therefore, the operating conditions of each stage are almost the same hydraulically, and the effective head of each stage is 2 as given by the following equation (5).
It is equal to the total effective head acting on the stage pump turbine divided into approximately two equal parts, and therefore the operating state of each stage follows a trajectory from □0 to □A.

What should be noted here is that, due to the above-mentioned problems with the hydraulic characteristics peculiar to single-speed reversible pump-turbines, the reference operating state point of each stage section is Low-head area with low hydraulic characteristics far away from the highest efficiency point (△η=0) (area where the unit rotational speed is larger than that at the highest efficiency point)
Therefore, it is necessary to select
When the guide vanes are controlled, the steady operating state locus of each step section also passes through a region with relatively low hydraulic characteristics, as shown by □0→□A in the figure.

As described above, in a normal two-stage pump turbine, the hydraulic performance of each stage is relatively low, and the current situation is that the turbine is forced to operate undesirably.

In such a case, the present invention reduces the outer diameter of the low-pressure stage runner as described above, so that the pressure at the runner outlet is relatively low compared to the high-pressure stage, thereby reducing cavitation and secondary water flow. Regarding the low-pressure stage section, which is a problem where problems such as By forming the runners to ensure stable, high-performance operation, the overall hydraulic performance of each stage is significantly improved.

A specific embodiment of the present invention will be described with reference to a single-speed reversible two-stage pump water turbine shown in FIG. In FIG. 1, the pump efficiency η under the single-stage hydraulic characteristics of the high-pressure stage runner 2 equipped with the movable guide vane 15
The unit rotational speed ratio between p and the turbine efficiency ηt at each maximum efficiency point is 1.04 to 1.16 (see formula (1)), and by reducing the outer diameter D of the blades of the low-pressure stage runner 3, The rotational speed N/√ ₀ is increased to shift the single-stage hydraulic characteristics of the low-pressure stage runner 3 to the low-head side compared to the high-pressure stage runner 2 (see formula (4)).

In this case, the relative relationship between the outer diameter D ₁ of the blades of the high-pressure stage runner 2 and the outer diameter D ₂ of the blades of the low-pressure stage runner 3 is expressed by _the equation (4 ) gives equation (6). However, the suffix 1 represents the high-pressure step portion, and the suffix 2 represents the low-pressure step portion.

In equation (6), the right-hand side is the unit rotational speed (N/√ ₀₂ ) of the high-pressure stage where the outer diameter of the runner blade remains the same as before when the unit rotational speed (N/√ 02 ) of the low-pressure stage is increased.
₀₁ ) as the denominator (N/√ ₀₂ )/
It corresponds to the reciprocal of (N/√ ₀₁ ), and the left side is the reduction ratio D ₂ / of the runner blade outer diameter D ₂ of the low-pressure stage with the runner blade outer diameter D 1 of the high-pressure stage as the denominator, which is the _same as before.
Equivalent to D ₁ .

In other words, Equation (6) shows that when increasing the unit rotational speed (N/√ ₀₂ ) of the low-pressure stage, the required reduction ratio D ₂ /D ₁ of the outer diameter D ₂ of the runner blade of the low-pressure stage is equal to the unit rotational speed. It shows that it is given by the reciprocal of the increase ratio. In this case, the runner blade outer diameter D ₁
However, the single-stage hydraulic characteristics of the high-pressure stage section are unchanged from conventional ones, and as mentioned above, the maximum efficiency unit rotational speed of the turbine is N/√
₀₁ is pump maximum efficiency unit rotational speed N/√ ₀₁
becomes smaller, and the relative ratio of N/√ ₀₂ and N/√ ₀₁ is 1.04 to 1.16 as given by equation (1) above.
The basic problem is that the water turbines are forced to operate at low performance. Therefore, considering the situation in the high-pressure stage, how can the reduction ratio of the outer diameter of the runner blade be reduced in the low-pressure stage?
The important point is whether to determine D ₂ /D ₁ and increase the unit rotational speed N/√ ₀₂ .

Therefore, in this embodiment, as mentioned above, in the high pressure stage section, the maximum efficiency unit rotational speed of the water turbine is N/√
₀₁ is pump maximum efficiency unit rotational speed N/√ ₀₁
The relative ratio of N/√ ₀₁ and N/√ ₀₁ is 1.04 to 1.16 times smaller. Therefore, as a target for increasing the unit rotational speed N/√ ₀₂ in the low-pressure stage, the increase ratio is higher than that of the high-pressure stage. The purpose is to specify the relative ratio to be equal to the above relative ratio, 1.04 to 1.16 times. On the other hand, the reduction ratio of the outer diameter _D2 of the low-pressure stage runner blade required to achieve this increase ratio of 1.04 to 1.16 times is calculated by adding the reciprocal of the increase ratio of 1.04 to 1.16 to the right side of equation (6), as described above. Since it can be obtained by substitution, the relative relationship between the outer diameters of the runner blades is ultimately given as in equation (7).

Equation (7) is the outer diameter of the runner blade of the runner in the low pressure stage.
D ₂ is 0.86 to 0.96 of the outer diameter of the runner blade of the high pressure stage D ₁
This means setting it to twice the range.

In this case, the hydraulic characteristics of each stage are shown in Figure 4 when the turbine is operating, and Figure 5 when the pump is operating.
They are shown in the figure. In other words, the outer diameter of the blades of the low pressure stage runner 3 is the outer diameter D ₁ of the blades of the high pressure stage runner 2.
The runners of each step are configured so that D ₂ is within a specific range defined by equation (7), and the water opening of the movable guide vanes 15 and 16 of each step is adjusted so that the effective head of each step is equal to the conventional one. As is clear from the trajectory of the turbine operation status of each stage shown in Fig. 4, especially in water turbine power generation operation where high-efficiency operation is strongly required, the high-pressure stage section In this case, operation is performed at the same performance level as before, but in the low pressure section, the entire performance of the water turbine is shifted to the low head side and it is operated at almost the same high head as before, so the low pressure section is operated at the standard operating state □ Since ideal high-performance operation can be performed from 0 to a small flow rate region, the overall hydraulic performance of each stage can be significantly improved as a result. In this way, in order to arrange runners with different blade outer diameters on each step and maintain the effective head of each step almost equally as in the conventional case where the outer diameter of the blades is the same, it is necessary to This can be done by adjusting and controlling the opening degree of the water port of the vane to such an opening degree relationship that the step portion with the smaller outer diameter of the blade has a relatively smaller opening.
In addition, in the case of pump operation, as shown in Figure 5,
Low-pressure stage runner 3 in which the pump performance on the high head side has shifted its entire performance characteristics to the low head side compared to the conventional case.
However, the high-pressure stage runner 2 has a large blade outer diameter and excellent high-head performance.
Since the recovery can be improved by the pump, the overall performance of each stage section can be operated at the same level of pump performance as in the conventional case.

As described above, the present invention provides a movable guide vane in each stepped section, and makes the outer diameter of the runner blade of the low-pressure stepped section smaller than that of the high-pressure stepped section within the above-mentioned specific range. The runners of each stage are constructed using a runner, and the water opening of the movable guide vane of each stage is adjusted and controlled so that the effective head of each stage is almost the same as before. High performance operation is possible.

[Brief explanation of the drawing]

FIG. 1 is a schematic sectional view showing the structure of a two-stage pump-turbine according to an embodiment of the present invention, FIG. FIG. 4 is an explanatory diagram showing the relationship between the water turbine characteristics of each stage in a two-stage pump-turbine according to an embodiment of the present invention. FIG. FIG. 2 is an explanatory diagram showing the relationship between pump characteristics of each stage in the two-stage pump turbine of the same embodiment. 1... Water turbine main shaft, 2... High pressure stage runner, 3...
Low pressure stage runner, 10... return passage, 11... spiral casing, 13... inlet valve, 14... penstock, 15... movable guide vane, 16... movable guide vane.

Claims (1)

[Claims] 1 In a single-speed reversible two-stage pump water turbine in which the high-pressure stage section and the low-pressure stage section are connected in series through a return passage, and the runners of each stage section are connected to a single main shaft of the water turbine to perform water turbine or pump operation, A movable guide vane that can change the opening of the water port is provided on the outer periphery of the runner in each stage, and the outer diameter of the runner blade in the low-pressure stage is 0.86 to 0.86 the outer diameter of the runner blade in the high-pressure stage.
A single-speed reversible two-stage pump turbine characterized by being set within the range of 0.96 times.