JP2015151986A - Fluid machinery - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a fluid machinery which has a relatively simple system configuration and in which the occurrence of surging can be avoided with small number of times of performance test to be carried out in advance.SOLUTION: The fluid machinery comprises a rotatable impeller 3, a variable inlet guide vane 2 for generating a swirl flow in fluid flowed into the impeller 3 and a controller 14 for controlling a rotational speed N of the impeller 3 and a rotation angle αof the variable inlet guide vane 2. The controller 14 stores data of relationship between a suction flow rate Q and a stage efficiency η of the fluid machinery and data of relationship between a theoretical head Hand a suction flow rate Q of the fluid machinery. Using the stored data and a swirl angle αof the fluid flowing into the impeller 3, the controller 14 obtains a flow rate of fluid generating surging as a surging occurrence limit flow rate Qand controls the rotational speed N of the impeller 3 and the rotation angle αof the variable inlet guide vane 2 on the basis of the obtained surging occurrence limit flow rate Q.

Description

  The present invention relates to a fluid machine, and more particularly, to a fluid machine having an impeller and a swirling flow generating device that generates a swirling flow in a fluid flowing into the impeller.

  A fluid machine such as a centrifugal type or a mixed flow type imparts energy to a working fluid such as a gas or a liquid with an impeller rotated by a driving machine, and generally requires reduction of power consumption and expansion of a stable operation range. In order to reduce power consumption and expand the stable operating range, among these fluid machines, especially in a single-stage fluid machine, rotational speed control that controls the rotational speed of the drive with an inverter, etc., and upstream of the impeller Control by a swirling flow generator such as an attached guide vane (hereinafter referred to as “IGV”) is often performed. In the control by the IGV (IGV control), the IGV opening (the rotation angle of the IGV, that is, the IGV blade attachment angle) is changed, and the working fluid is imparted to the working fluid by the IGV in a direction that reduces the work of the impeller.

  If the operating flow rate of the fluid machine is reduced by reducing the rotational speed or the IGV opening, the difference between the blade inlet angle and the inflow angle of each component of the fluid machine such as the impeller and diffuser becomes large. The flow is separated from the blade surface at a certain flow rate due to the increase in deceleration of the internal flow velocity in the flow direction. When flow separation occurs, the pressure loss increases, and eventually a so-called stall occurs in which the pressure does not increase even if the flow rate is reduced. When the fluid machine stalls, that is, when a region of dH / dQ ≦ 0 appears on the characteristic curve of flow rate Q-head (pressure) rise H of the fluid machine, minute fluctuations in the flow rate and pressure inside the fluid machine are self-excited. Surging occurs that causes vibration and vigorously vibrates the pressure and flow rate of the entire system. When surging occurs, not only stable operation becomes impossible, but the equipment may be damaged.

  For this reason, in a fluid machine, it is necessary to perform operation control so that surging does not occur. When rotational speed control or IGV control is used, the surging generation limit flow rate, that is, the flow point at which dH / dQ = 0 is different for each rotational speed and IGV setting condition, so surging occurs at each rotational speed and each setting condition. It is necessary to determine the critical flow rate, and it is necessary to perform operation control using a control method that performs such determination.

  In the case of using IGV control, the simplest method for determining the surging occurrence limit flow rate is to perform a performance test in advance for a large number of IGV openings, and obtain the surging occurrence limit flow rate at each IGV opening, It is a technique to construct a database that collects these data. The characteristic change of the fluid machine when the rotation speed of the impeller changes, for example, as described in Non-Patent Document 1, the rotation speed similarity law of the fluid machine represented by the following equations (1) to (3):

(However, Q, H, and L represent the suction volume flow rate, the head (pressure), and the shaft power when the rotational speed of the impeller is N, respectively, and Q ′, H ′, and L ′ represent the blades, respectively. This represents the suction volume flow, the head, and the shaft power when the rotational speed of the vehicle is N ′. Therefore, if a database that collects data when the IGV opening is changed is constructed, it is sufficient to determine the surging generation limit flow rate.

  Many techniques have been proposed in the past for detecting surging by avoiding surging. For example, Patent Literature 1 describes a technique for detecting the beginning of surging by increasing the pressure.

JP 63-161362 A

Edited by Turbomachinery Association, “Turbo Pump”, Nihon Kogyo Shuppan, 2007, pp. 19-20

  Even with the conventional method as described above, it is possible to avoid the occurrence of surging. However, the system configuration is complicated because there is a problem that it is necessary to perform a large number of performance tests in advance to determine the surging limit flow rate, and a high-response sensor is required to detect the beginning of surging. There was a problem of becoming expensive.

  SUMMARY OF THE INVENTION An object of the present invention is to provide a fluid machine in which the number of performance tests to be performed in advance is small and surging can be avoided with a relatively simple system configuration.

  In order to solve the above problems, a fluid machine according to the present invention includes a rotatable impeller, a variable inlet guide vane that is rotatable and generates a swirling flow in the fluid flowing into the impeller, and the rotation of the impeller. And a controller for controlling the speed and the rotation angle of the variable inlet guide vane. The controller stores data on the relationship between the suction flow rate of the fluid machine and the stage efficiency, and data on the relationship between the theoretical head of the fluid machine and the suction flow rate, and the data flows into the impeller. From the swirl angle of the fluid, the flow rate of the fluid that generates surging is obtained as a surging generation limit flow rate, and based on the calculated surging generation limit flow rate, the rotational speed of the impeller and the rotation angle of the variable inlet guide vane It is characterized by controlling.

  According to the present invention, it is possible to provide a fluid machine in which the number of performance tests to be performed in advance is small and surging can be avoided with a relatively simple system configuration.

The longitudinal cross-sectional view of the fluid machine by Example 1 of this invention. An actual measurement example of the characteristic curve of the centrifugal single-stage fluid machine A. An actual measurement example of the characteristic curve of the centrifugal single-stage fluid machine B. The figure which shows the characteristic curve of the stage efficiency (eta) of the centrifugal single-stage fluid machine. The figure which shows the characteristic curve of the impeller efficiency (eta) _imp of the centrifugal single-stage fluid machine B. FIG. The figure which shows the absolute velocity vector in the position of the diffuser of a fluid machine. The figure which shows the speed triangle in the exit of the impeller of a fluid machine. The figure which shows the absolute speed vector in the downstream of IGV of a fluid machine, and the speed triangle in the inlet of an impeller. The schematic diagram of the performance characteristic curve of a fluid machine. The figure which shows typically the relationship between IGV rotation angle and the pre-swirl angle (alpha) ₁ of the actual fluid flow. 1 is a configuration diagram of a fluid machine control system in Embodiment 1. FIG. 3 is a flowchart of a method for deriving a surging generation limit flow rate of the fluid machine in the first embodiment. The block diagram of the control system of the fluid machine in Example 2 of this invention.

  The fluid machine according to the present invention includes an impeller, a variable inlet guide vane (IGV) that is a swirling flow generating device that generates a swirling flow in the fluid flowing into the impeller, and the rotational speed of the impeller and the rotation of the variable inlet guide vane. And a controller that controls the angle, and the surging generation limit flow rate, which is the flow rate of surging, is generated with a relatively simple system configuration while minimizing the number of performance tests to be performed in advance. Can be calculated. As a result, the fluid machine according to the present invention can avoid the occurrence of surging and improve the operation reliability.

  Hereinafter, a fluid machine according to an embodiment of the present invention will be described with reference to the drawings, taking a single-stage fluid machine having only one impeller as an example.

  1 is a longitudinal sectional view of a fluid machine according to a first embodiment of the present invention. In FIG. 1, among the fluid machines, the upper side of the rotating shaft 6 of the impeller 3 is shown, and the lower side is omitted. In this embodiment, a single-stage centrifugal turbo compressor will be described as an example of the fluid machine. However, the fluid machine according to the present invention is also applicable to other types of fluid machines. Can do.

  The fluid machine according to the present embodiment imparts an arbitrary swivel angle to the impeller 3 that rotates to boost the fluid, the rotary shaft 6 on which the impeller 3 is installed, and the fluid that is rotatably installed and flows into the impeller 3. IGV2 that generates a swirling flow, a drive unit 1 that rotates the IGV2, a circular blade row diffuser 4 that is provided on the downstream side of the impeller 3 and converts the dynamic pressure of the fluid flowing in from the outlet of the impeller 3 to static pressure, And a discharge scroll 5 connected to the diffuser 4. The IGV 2 is a blade row provided at the inlet of the working fluid of the fluid machine and provided radially with respect to the rotating shaft 6. The impeller 3 is rotationally driven by the rotation of the rotating shaft 6 by the action of a driving machine (not shown). The diffuser 4 may be a vaneless diffuser without a wing. Further, the discharge scroll 5 may be a discharge channel having another form similar to the scroll. Each component of these fluid machines is stored in the casing 7 and constitutes an internal flow path.

  The working fluid is sucked in from the suction flow path 8 of the fluid machine, flows into the impeller 3 in a state where an arbitrary turning angle is given by the IGV 2, is increased in pressure, decelerated by the diffuser 4, and then discharged by the discharge scroll 5. It is led to the flow path 9.

  The performance characteristics of the fluid machine according to the first embodiment of the present invention will be described with reference to FIGS. 2A and 2B. 2A and 2B are two types of single-stage centrifugal fluid machines (centrifugal single-stage fluid machine A and centrifugal single-stage fluid machine) having the same configuration as in FIG. It is a figure which shows the actual measurement example of the characteristic curve of suction flow Q-head rise H, and the characteristic curve of suction flow Q-efficiency (stage efficiency) (eta) about B). The stage efficiency is the efficiency calculated from the head calculated from the temperature and pressure at the inlet of the IGV 2 and the outlet of the discharge scroll 5 and the theoretical head (shaft power). FIG. 2A shows a characteristic curve for the centrifugal single-stage fluid A, and FIG. 2B shows a characteristic curve for the centrifugal single-stage fluid machine B. As shown in FIGS. 2A and 2B, for the centrifugal single-stage fluid A, a performance test is performed when the IGV opening is 100%, 35%, and 15%, and for the centrifugal single-stage fluid B, A performance test was conducted when the IGV opening was 100%, 10%, and 5%, and characteristic curves were obtained.

  As can be seen from FIG. 2A and FIG. 2B, it was found from the results of the performance test that we conducted that the efficiency characteristics have the following characteristics. In other words, the two types of single-stage centrifugal fluid machines (centrifugal single-stage fluid machine A and centrifugal single-stage fluid machine B) that have been subjected to performance tests have different IGV opening degrees despite their completely different designs. It has been found that the Q-η characteristic curve in the vicinity of the surging limit flow rate on the low flow rate side can be expressed almost completely by a single curve (dotted line in FIGS. 2A and 2B) that does not depend on the IGV opening. The surging generation limit flow rate is a flow rate at which the slope of the characteristic curve of suction flow rate Q-head rise H becomes zero, that is, a flow rate at which dH / dQ = 0.

  Thus, in a single-stage fluid machine equipped with an IGV, the Q-η characteristic curve (relationship between the suction flow rate Q and the stage efficiency η) near the surging limit flow rate on the low flow rate side is not dependent on the IGV opening. The reason why it can be expressed by the curve of the book is considered as follows.

3A is a diagram showing a characteristic curve of the stage efficiency η of the centrifugal single-stage fluid machine B shown in FIG. 2B, and FIG. 3B shows a characteristic curve of the impeller efficiency η _imp of the centrifugal single-stage fluid machine B. FIG. In each of the IGV opening series, the vicinity of the flow rate on the lowest flow rate side is the surging generation limit flow rate at which dH / dQ = 0. The impeller efficiency η _imp is an efficiency calculated from the head calculated from the temperature and pressure at the inlet and outlet of the impeller 3 and the theoretical head (shaft power), and is the efficiency of the impeller 3 alone. It is. In addition, the scale of the vertical axis | shaft of FIG. 3A and FIG. 3B is the same.

As shown in FIG. 3A, the stage efficiency η varies greatly depending on the IGV opening. On the other hand, as shown by the dotted circle in FIG. 3B, the difference in the impeller efficiency η _imp near the surging generation limit flow rate at different IGV opening is near the surging generation limit flow at different IGV opening shown in FIG. 3A. Compared with the difference in the stage efficiency η, it is very small. Therefore, on the low flow rate side, the loss due to the diffuser 4 located on the downstream side of the impeller 3 is dominant, and the stage efficiency η is considered to change according to the efficiency of the diffuser 4.

  4A to 4C are diagrams showing velocity vectors at positions of the IGV 2, the impeller 3, and the diffuser 4 of the fluid machine. FIG. 4A is a diagram illustrating an absolute velocity vector at the position of the diffuser 4. FIG. 4B is a diagram showing a speed triangle at the exit of the impeller 3. FIG. 4C is a diagram showing an absolute velocity vector downstream of the IGV 2 and a velocity triangle at the inlet of the impeller 3. In addition, the speed triangle at the position of IGV2 indicates that the rotation angle of IGV2 (IGV rotation angle) is 0 ° (that is, the IGV opening is 100%) and that the IGV rotation angle is not 0 ° (that is, the IGV opening). Is not 100%).

4A to 4C, α _IGV is an IGV rotation angle (the angle formed by the rotation axis direction and the direction toward the rear edge of IGV 2 and the same direction as the rotation direction of impeller 3 is positive), and C is a fluid , Cm is the meridional component of the absolute velocity of the fluid, Cu is the circumferential component of the absolute velocity of the fluid, U is the circumferential velocity of the impeller 3, W is the relative velocity of the impeller 3, and α is the absolute velocity of the fluid The flow angle (the angle formed by the rotation axis direction or radial direction and the absolute velocity vector, the same direction as the rotation direction of the impeller 3 is positive), β is the relative flow angle of the fluid, and β _b is the impeller 3 blade angles (the angle formed by the circumferential direction and the tangent line drawn in the direction along the blades in each blade part, the same direction as the rotation direction of the impeller 3 being positive) “1” is the value at the entrance of the impeller 3, “2” is the value at the exit of the impeller 3, and “3” "Is a value at the inlet of the diffuser 4, and" 4 "is a value at the outlet of the diffuser 4.

As can be seen from FIGS. 4A to 4C, the absolute flow angle α ₁ (pre-turning angle α ₁ ) of the fluid flowing into the impeller 3 varies depending on the IGV rotation angle α _IGV , and the blades according to the value of the pre-turning angle α _1. It is predicted that the shape of the Q-η _imp characteristic curve of the vehicle 3 will change greatly. This is the theoretical head H _{th of the} impeller represented by the following formula (4) (head without loss, efficiency 100%),

(Where g represents gravitational acceleration) and the impeller 3 incident angle (= β _1b −β ₁ ) and impeller 3 reduction ratio (= W ₁ / W ₂ ) This is because it largely changes depending on the value of the pre-swivel angle α ₁ at the entrance of No. 3.

However, considering an ideal state where there is no pressure loss at the IGV portion and the working fluid is an incompressible fluid, even if α ₁ changes, the fluid in the meridional direction at the outlet of the impeller 3 since the absolute change in velocity Cm ₂ is not, if the same suction flow Q, the shape of the velocity triangle of the outlet of the impeller 3 regardless of the alpha ₁ value, a coincident (in fact, the working fluid Is a compressible fluid, increasing α _IGV to increase α ₁ causes pressure loss in the IGV section and changes the density of the working fluid at the inlet of the impeller 3, so Cm ₂ is strictly Are not equal, but the amount of change is small, so the above ideal state can be assumed). Therefore, the diffuser incident angle (which is a factor that greatly affects the shape of the efficiency characteristic curve of the diffuser 4 (curve with the horizontal axis representing the flow rate and the vertical axis representing the ratio of the head loss occurring at each diffuser to the theoretical head). = Β _3b −α ₃ ) is also determined by only the value of the suction flow rate Q, regardless of the value of α ₁ . Since the IGV rotation angle α _IGV and the working fluid pre-swivel angle α ₁ have a one-to-one correspondence, as a result, the shape of the efficiency characteristic curve of the diffuser 4 is substantially the same regardless of the value of α _IGV. It is thought that it becomes.

As described above, the impeller efficiency η _{imp in the} vicinity of the surging limit flow rate is substantially the same value regardless of the IGV opening, and the shape of the efficiency characteristic curve of the diffuser 4 that is dominant to the performance on the low flow rate side. Is considered to be substantially the same regardless of the value of α _IGV , and as a result, the Q-η characteristic curve in the vicinity of the surging limit flow rate can be expressed as a single curve independent of the IGV opening. .

  Using this knowledge, it is possible to calculate the surging generation limit flow rate of a fluid machine with a relatively simple system configuration while minimizing the number of performance tests to be performed in advance. The method will be described below.

From the above equations (4) and FIG. _{4A-4C, H th} are guided to be linearly increased with the decrease of the suction flow rate Q. When the pre-swivel angle α ₁ is not given to the fluid, the second term U ₁ Cu ₁ / g on the right side of Equation (4) becomes 0, and H _th is determined only by the velocity triangle at the exit of the impeller 3. . On the other hand, when the pre-swivel angle α ₁ is given to the fluid, H _{th at the} same flow rate becomes smaller than when the pre-swivel angle α ₁ is not given, but Cu ₁ becomes smaller as the flow rate is reduced. , the difference between the H _th when no given pre-rotation angles alpha ₁ gradually decreases. In each flow point, since the value obtained by subtracting the pressure loss of each unit from H _th is the actual fluid machine head rise H, the stage efficiency η of the fluid machine, the following equation (5)

It is represented by

FIG. 5 is a schematic diagram of the performance characteristic curve of the fluid machine, and shows the head rise H and the stage efficiency η with respect to the suction flow rate Q. As described above, a single Q-η characteristic curve (FIG. 1) that does not depend on the IGV opening near the surging generation limit flow rate is required as long as a performance test is performed in advance by changing α _IGV in two or three ways. Η ′ (Q)) in FIG. Therefore, in addition to this, if the flow rate change characteristic of the theoretical head H _th according to the value of α _IGV is known, since H = η′H _th , the surging occurrence limit is satisfied for all IGV opening conditions. It is possible to derive the shape of the QH characteristic curve near the flow rate.

In _obtaining the flow rate change characteristic of the theoretical head _Hth , the first term on the right side of the equation (4) representing _Hth can be obtained, for example, as follows. From the result of the performance test performed in advance, the slip coefficient ξ (an index indicating how much the flow flows along the blade at the exit of the impeller 3) represented by the following formula (6),

(However, Cu _2∞ represents the circumferential component of the absolute velocity of the fluid when the flow completely flows out along the blade.) If the working fluid is an incompressible fluid,

Thus, Cu ₂ can be obtained only from the slip coefficient ξ, the representative dimension values of the impeller 3 (for example, the outlet diameter, the outlet width, and the outlet angle) and the suction flow rate Q. When the working fluid is a compressive fluid, the density of the fluid changes between the inlet and outlet of the impeller 3, so that the volumetric flow rate at the outlet of the impeller 3 is not the same as the suction flow rate Q. By carrying out, Cu ₂ can be obtained. Here, since the value of ξ is hardly affected by the change in α _IGV , the value of ξ is calculated from the performance test result when the IGV opening is 100% so that the calculated value of H _th matches the measured value. do it.

On the other hand, for the second term on the right side of the equation (4), some method for deriving the value of the pre-turn angle α ₁ shown in FIGS. This is because the IGV rotation angle α _IGV and the actual fluid flow pre-swirl angle α ₁ do not always match. In this embodiment, the relationship between α _IGV and α ₁ is derived from the results of performance tests in which α _IGV is changed in two to three ways. As described above, this performance test is performed in advance in order to obtain a Q-η characteristic curve (η ′ (Q) in FIG. 5) independent of the IGV opening.

FIG. 6 is a diagram schematically showing the relationship between the IGV rotation angle α _IGV derived from the result of the performance test in which α _IGV is changed in two to three ways and the actual swirl angle α ₁ of the actual fluid flow. . Theoretical head H calculated using the results of several performance tests performed in advance by changing the IGV opening (that is, the IGV rotation angle α _IGV ) and the slip coefficient ξ obtained by the above equation (6), for example. The value of α ₁ is obtained so that _th matches. At this time, the following equation (8)

And gradually changing the ₁ alpha into Equation (4), the alpha ₁ when calculated H _th is consistent with the measured values, the actual pre-rotation angles. For each IGV opening (ie, IGV rotation angle α _IGV ) for which the preliminary test has been performed, the value of α ₁ is obtained, these values are plotted on a graph, and interpolation between these plots and the point (0, 0) is performed. If the interpolation function is calculated, the relationship α ₁ (α _IGV ) between the IGV rotation angle α _IGV and the actual fluid flow pre-swivel angle α ₁ can be derived, and the α _IGV -α ₁ curve shown in FIG. Can be obtained. From the α _IGV -α ₁ curve (interpolation function α ₁ (α _IGV )) shown in FIG. 6, the pre-turn angle α _{1 at} all IGV opening degrees is obtained.

In FIG. 6, a broken line representing α ₁ = α _IGV is shown for reference. When the fluid flows along IGV2, α ₁ = α _IGV .

By _{obtaining the} α _IGV -α ₁ curve from the results of the preliminary test, the shape of the QH characteristic curve in the vicinity of the surging limit flow rate at all IGV openings (ie, all IGV rotation angles α _IGV ) is derived. can do. And the surging generation | occurrence | production limit flow volume of a fluid machine can be calculated | required by calculating | requiring the suction flow volume Q used as dH / dQ = 0 in the calculated | required QH characteristic curve.

  FIG. 7 is a configuration diagram of a fluid machine control system in the present embodiment. As illustrated in FIG. 7, the fluid machine control system includes the fluid machine illustrated in FIG. 1, a measuring device installed in the fluid machine, and a controller 14 that controls the fluid machine.

In the suction flow path 8 of the fluid machine, a suction flow rate Q measuring device 11a, a suction pressure Ps measuring device 11b, and a suction temperature Ts measuring device 11c are installed. An IGV rotation angle α _IGV measuring device 11d is installed in the driving machine 1 of the IGV2. On the rotating shaft 6 of the impeller 3, a measuring device 11e for the rotational speed N of the impeller 3 is installed. In the discharge channel 9, a measuring device 11f for the discharge pressure Pd is installed. These measuring devices 11a to 11f are connected to the controller 14 via lines 12a to 12f.

The controller 14 sends a control signal for adjusting the IGV rotation angle α _IGV to the driver 1 of the IGV 2 via the line 13a. In the IGV 2, the IGV opening (that is, the IGV rotation angle α _IGV ) is changed by the drive unit 1. Further, the controller 14 sends a control signal for adjusting the rotation speed N to the driving machine 10 of the impeller 3 through the line 13b. The impeller 3 is rotationally driven when the rotating shaft 6 is rotated by the driving machine 10. Although not shown in FIG. 7, when a device such as a flow rate adjustment valve is incorporated in the pipe, the controller 14 can also transmit a signal for adjusting these devices to the device. .

The controller 14 includes various arithmetic devices and a storage device that stores various data. Examples of the calculation device include a device for obtaining a change in performance of the fluid machine when the rotational speed N of the impeller represented by the equations (1) to (3) is changed, and Cu ₂ when the working fluid has compressibility. There are a device for performing a simple iterative operation for obtaining the value, a device for performing other operations necessary for determining the surging limit flow rate, and the like. As an example of data stored in the storage device, η ′ (Q), a slip coefficient ξ obtained from a result of a performance test performed in advance, an H obtained from a performance test result when the IGV opening is 100%. _{There is} a flow rate change characteristic and a gradient of _th , data of an interpolation function α ₁ (α _IGV ), main dimension values of a target fluid machine, and other data necessary for determining a surging limit flow rate.

In the present embodiment, the surging generation limit flow rate of the fluid machine is obtained and the operation of the fluid machine is controlled as follows. In the following description, in order to obtain a desired head H and flow rate Q, a fluid machine that has been operated at a certain rotational speed and IGV rotational angle is operated at a different rotational speed N ′ and IGV rotational angle α _IGV ′. Consider the case.

  FIG. 8 is a flowchart of a method of deriving the surging generation limit flow rate of the fluid machine in the present embodiment.

  In S10, the rotational speed N ′ of the impeller 3 is acquired by the measuring device 11e (FIG. 7).

  In S12, N ′ acquired in S10 is compared with the rotational speed N of the impeller 3 in the performance test performed in advance, and the rotational speed ratio N ′ / N is calculated.

  In S14, the rotational speed ratio N '/ N and the formula (9) obtained by modifying the formula (1).

From this, the operating point (flow rate Q ') at the rotational speed N' that satisfies the same operating conditions (that is, the efficiency becomes equal) as the operating point (flow rate Q) at the rotational speed N is obtained.

  In S16, if this flow rate Q ′ is substituted for Q of the Q-η characteristic curve (η ′ (Q) in FIG. 5) obtained in advance, it does not depend on the IGV opening at the rotational speed N ′. One Q-η characteristic curve η ′ (Q ′)

Get.

In S18, it obtains the flow rate change characteristic of the theoretical head _{H th} 'IGV rotation angle is alpha _IGV' rotational speed N in the conditions, i.e. the theoretical head characteristics _{H th} the '(Q'). Specifically, the circumferential component Cu ₂ ′ (Q ′) of the absolute velocity of the fluid at the outlet of the impeller 3 under this operating condition is calculated from the value of ξ obtained in advance and Equation (7). Ask. Furthermore, from the interpolation function α ₁ (α _IGV ) (FIG. 6) and Equation (8) obtained in advance, the circumferential component Cu ₁ of the absolute velocity of the fluid at the inlet of the impeller 3 under these operating conditions. Find '(Q'). The IGV rotation angle α _IGV ′ is measured by the measuring device 11d. H _th ′ (Q ′) is obtained from Cu ₂ ′ (Q ′), Cu ₁ ′ (Q ′) and Equation (4) obtained as described above.

In S20, when η ′ (Q ′) obtained in S16 and H _th ′ (Q ′) obtained in S18 are substituted into Equation (5), the flow rate change of the head in the vicinity of the surging generation limit flow rate under the present operating conditions. The characteristic H ′ (Q ′) is expressed by the following equation (11)

Can be obtained.

In S22, Equation (11) is differentiated by Q ′, and a flow rate Q ′ at which dH ′ (Q ′) / dQ ′ = 0 is obtained as a surging generation limit flow rate Q _surge under the present operating conditions.

As described above, in the fluid machine in the present embodiment, the controller 14 calculates the flow rate Q _{surge at} which dH ′ (Q ′) / dQ ′ = 0, and repeatedly stores and updates the calculated flow rate Q _surge . The controller 14 calculates the head H and the flow rate Q calculated based on the measurement results of the flow rate Q, the suction pressure Ps, the suction temperature Ts, and the discharge pressure Pd, and the flow rate Q is a surging occurrence limit. The rotation speed N of the impeller 3, the IGV rotation angle α _IGV , and the flow rate adjustment valve are controlled so as not to fall below the flow rate Q _surge .

2 that the maximum efficiency point is located near the line of η (Q) ′ when α _IGV ≠ 0 ° (when the IGV opening is not 100%). Therefore, when α _IGV is set large (when the IGV opening is reduced), the rotational speed N of the impeller 3, the IGV rotation angle α _IGV , and the flow rate aiming at the flow rate on the line η ′ (Q). It is preferable to control the adjusting valve so as to match the desired head H.

  FIG. 9 is a configuration diagram of a control system for a fluid machine in Embodiment 2 of the present invention. 9, the same reference numerals as those in FIG. 7 denote the same or common components as those in FIG. 7, and description of these components is omitted.

The fluid machine control system according to the present embodiment includes a detection device 15 having a pre-turn angle α ₁ between the IGV 2 and the impeller 3, and the detection device 15 is connected to the controller 14 via a line 15 a. The point is different from the control system (FIG. 7) of the fluid machine according to the first embodiment. By providing the detection device 15 for the pre-turning angle α ₁ , the surging occurrence limit flow rate Q _surge is obtained in the same manner as in the first embodiment without obtaining the interpolation function α ₁ (α _IGV ) shown in FIG. 6 in advance. It is possible. The surging generation limit flow rate Q _surge can be obtained by the method according to the flowchart shown in FIG. However, in S18, the circumferential component Cu ₁ ′ (Q ′) of the absolute velocity of the fluid at the inlet of the impeller 3 is obtained from the pre-turning angle α ₁ detected by the detection device 15 and the equation (8).

The pre-rotation angle alpha ₁ of the detector 15, for example, may be used a lightweight metal pieces small such changes direction completely follow the flow of fluid downstream of IGV2. Such metal piece, mounted on a rotatable shaft, inserted into the interior of the pipe, by measuring the rotational angle of the metal strip at an angle detector detects the pre-rotation angle alpha _1.

Since the fluid machine control system according to the present embodiment includes the detection device 15, the configuration is slightly complicated as compared with the fluid machine control system according to the first embodiment, but the pre-turn angle α ₁ can be directly obtained. . Therefore, than the fluid control system of the machine of the first embodiment precision how interpolation depends upon obtaining the pre-rotation angle alpha _1, it is possible to more accurately determine the surging occurrence limit flow rate Q _surge.

  In addition, this invention is not limited to said Example, Various modifications are included. For example, the above-described embodiments are described in detail for easy understanding of the present invention, and the present invention is not necessarily limited to an aspect including all the configurations described.

DESCRIPTION OF SYMBOLS 1 ... IGV drive machine, 2 ... IGV (variable inlet guide vane), 3 ... Impeller, 4 ... Diffuser, 5 ... Discharge scroll, 6 ... Rotating shaft of impeller, 7 ... Casing, 8 ... Suction flow path, 9 DESCRIPTION OF SYMBOLS ... Discharge flow path, 10 ... Impeller drive machine, 11a ... Suction flow rate Q measuring device, 11b ... Suction pressure Ps measuring device, 11c ... Suction temperature Ts measuring device, 11d ... IGV rotation angle α _IGV measuring device , 11e: a measuring device for the rotational speed N of the impeller, 11f: a measuring device for the discharge pressure Pd, 12a to 12f, 13a, 13b, 15a ... a circuit, 14 ... a controller, 15 ... a detecting device for a pre-turning angle.

Claims (3)

A rotatable impeller,
A variable inlet guide vane that is rotatable and generates a swirling flow in the fluid flowing into the impeller;
A controller for controlling the rotational speed of the impeller and the rotational angle of the variable inlet guide vane,
The controller is
Stores data on the relationship between the suction flow rate of the fluid machine and the stage efficiency, and data on the relationship between the theoretical head of the fluid machine and the suction flow rate,
From these data and the swirl angle of the fluid flowing into the impeller, the flow rate of the fluid at which surging occurs is determined as the surging generation limit flow rate,
Based on the calculated surging limit flow rate, the rotational speed of the impeller and the rotational angle of the variable inlet guide vane are controlled.
A fluid machine characterized by that. The controller is
Further storing data of the relationship between the rotation angle of the variable inlet guide vane and the turning angle of the fluid flowing into the impeller,
The fluid machine according to claim 1, wherein a swirl angle of the fluid flowing into the impeller is obtained from the data, and the surging generation limit flow rate is obtained using the obtained swirl angle of the fluid. Further comprising a detection device for a swirl angle of the fluid flowing into the impeller between the variable inlet guide vane and the impeller,
The fluid machine according to claim 1, wherein the controller obtains the surging generation limit flow rate using a swirl angle of the fluid detected by the detection device.

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