JPH0339202B2 - - Google Patents
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- Publication number
- JPH0339202B2 JPH0339202B2 JP57013251A JP1325182A JPH0339202B2 JP H0339202 B2 JPH0339202 B2 JP H0339202B2 JP 57013251 A JP57013251 A JP 57013251A JP 1325182 A JP1325182 A JP 1325182A JP H0339202 B2 JPH0339202 B2 JP H0339202B2
- Authority
- JP
- Japan
- Prior art keywords
- water supply
- steam
- supply pump
- turbine
- flow rate
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Expired - Lifetime
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- XLYOFNOQVPJJNP-UHFFFAOYSA-N water Substances O XLYOFNOQVPJJNP-UHFFFAOYSA-N 0.000 claims description 68
- 230000008878 coupling Effects 0.000 claims description 11
- 238000010168 coupling process Methods 0.000 claims description 11
- 238000005859 coupling reaction Methods 0.000 claims description 11
- 239000012530 fluid Substances 0.000 claims description 6
- 238000000034 method Methods 0.000 claims description 4
- 230000001105 regulatory effect Effects 0.000 description 27
- 238000010586 diagram Methods 0.000 description 8
- 238000012937 correction Methods 0.000 description 7
- 230000007423 decrease Effects 0.000 description 2
- 238000010248 power generation Methods 0.000 description 2
- 230000005540 biological transmission Effects 0.000 description 1
- 230000001276 controlling effect Effects 0.000 description 1
- 230000003247 decreasing effect Effects 0.000 description 1
- 238000001514 detection method Methods 0.000 description 1
- 230000004043 responsiveness Effects 0.000 description 1
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- Control Of Positive-Displacement Pumps (AREA)
- Flow Control (AREA)
Description
【発明の詳細な説明】
本発明は、制御特性の良い給水ポンプの回転数
制御方式に関するものである。DETAILED DESCRIPTION OF THE INVENTION The present invention relates to a water pump rotation speed control system with good control characteristics.
火力発電設備に於けるボイラ給水ポンプあるい
は原子力発電設備に於ける原子炉給水ポンプは蒸
気タービンによつて駆動されることが多い。 Boiler feed water pumps in thermal power generation equipment or reactor feed water pumps in nuclear power generation equipment are often driven by steam turbines.
これらポンプを駆動する蒸気タービンは、ボイ
ラあるいは原子炉の制御の一環として、回転数を
制御され、これにより給水流量が制御される。即
ち、ボイラあるいは原子炉の給水制御装置の信号
により、蒸気タービンの回転数を変化させて所望
の給水量を得るものである。 The rotation speed of the steam turbines that drive these pumps is controlled as part of boiler or nuclear reactor control, and thereby the feed water flow rate is controlled. That is, a desired water supply amount is obtained by changing the rotational speed of a steam turbine based on a signal from a water supply control device of a boiler or a nuclear reactor.
しかし、蒸気タービンの回転数制御範囲が広い
こと、および蒸気タービン作動の蒸気源を主ター
ビンの抽気としていることにより、蒸気圧の変
化、及び給水ポンプの吐出流量が非線形である等
の理由によりその制御は困難である。 However, because the speed control range of the steam turbine is wide and the steam source for operating the steam turbine is extracted air from the main turbine, changes in steam pressure and the discharge flow rate of the feedwater pump are nonlinear, etc. Control is difficult.
特に、後述するように、低流量域では制御が不
安定となり、速度制御による運転が出来ないとい
う制約がある。このため、低流量域では、給水調
整弁による制御、あるいはモータ駆動ポンプに切
替えていた。 In particular, as will be described later, control becomes unstable in a low flow rate region, and there is a restriction that operation by speed control is not possible. For this reason, in low flow areas, control has been switched to a water supply regulating valve or a motor-driven pump.
このような給水調整弁による制御は、弁の絞り
ロスによる効率低下の問題があり、ポンプ切替運
転は切替時間の制約の問題があつた。 Control using such a water supply regulating valve has the problem of decreased efficiency due to throttling loss of the valve, and pump switching operation has the problem of restrictions on switching time.
また、従来の機械油圧式制御装置では、レバー
系の遊びのため不感帯が大きく、回転数検出器と
しての調速機の検出範囲がせまいこと、また、そ
の特性も非線形であることから、ボイラ給水制御
を良好に行うことができないなどの欠点があつ
た。 In addition, with conventional mechanical-hydraulic control devices, the dead zone is large due to play in the lever system, the detection range of the speed governor as a rotation speed detector is narrow, and its characteristics are nonlinear, so boiler water supply There were drawbacks such as inability to perform good control.
本発明は、従来の機械油圧式制御装置を電子油
圧式制御装置に置換えることにより、駆動蒸気源
の圧力変動の影響を補正して制御特性を改善する
とともに、低流量域では高ゲインとなる給水対速
度特性を改善し、給水ポンプ駆動用タービンの制
御に適した装置を具備した給水ポンプの回転数制
御方式を提供することを目的とする。 By replacing the conventional mechanical-hydraulic control device with an electro-hydraulic control device, the present invention improves control characteristics by correcting the influence of pressure fluctuations of the driving steam source, and achieves high gain in the low flow rate region. It is an object of the present invention to provide a water supply pump rotation speed control system that improves the water supply versus speed characteristics and is equipped with a device suitable for controlling a turbine for driving the water supply pump.
本発明の要旨は、給水要求値およびポンプ吐出
圧力に対応して給水ポンプの速度指令値を正確に
求めるとともに、必要な場合にはさらに、タービ
ン駆動蒸気条件をフイードバツクし、タービン加
減弁開度を補正するようにした点にある。 The gist of the present invention is to accurately determine the speed command value of the feed water pump in accordance with the water supply demand value and the pump discharge pressure, and, if necessary, to feed back the turbine drive steam conditions and adjust the turbine control valve opening. The problem lies in the fact that it has been corrected.
このような構成とすることにより、本発明は、
従来速度制御が出来なかつた低流量域を含めた広
い速度範囲にわたつて、給水ポンプを適正に制御
し、安定な給水制御を応答性良く行わせるように
したものである。 With such a configuration, the present invention
This system appropriately controls the water supply pump over a wide speed range, including the low flow rate range where speed control could not be performed conventionally, and performs stable water supply control with good responsiveness.
以下本発明の一実施例を、図面を参照して説明
する。 An embodiment of the present invention will be described below with reference to the drawings.
第１図は本発明の一実施例のブロツク図であ
る。図に於いて、蒸気タービン１は、給水ポンプ
２に直結され、これを駆動するものであり、その
軸端には歯車５が設けられている。蒸気は、高圧
加減弁３または／及び低圧加減弁４を通つて、蒸
気タービン１に流入し、タービンを回転させ、直
結の給水ポンプ２を駆動させる。 FIG. 1 is a block diagram of one embodiment of the present invention. In the figure, a steam turbine 1 is directly connected to and drives a water supply pump 2, and a gear 5 is provided at the end of the shaft. Steam flows into the steam turbine 1 through the high pressure regulating valve 3 and/or the low pressure regulating valve 4, rotates the turbine, and drives the directly connected feed water pump 2.
タービン１の回転数は、歯車５に対向配置され
た速度検出器６，７によりパルス例として検出さ
れ、速度信号変換器８，９で演算信号に変換され
る。真値選択器１０は、２つの速度信号のうち真
値とされる値を選択する。 The rotation speed of the turbine 1 is detected as an example of a pulse by speed detectors 6 and 7 disposed opposite to the gear 5, and is converted into a calculation signal by speed signal converters 8 and 9. The true value selector 10 selects the true value from among the two speed signals.
前述の真値選択器１０は、２信号の差が許容値
以内であれば、高値を真値とし、差が許容値以上
であれば変化率の小さい方を真値とする処理が行
なわれる。 The aforementioned true value selector 10 performs a process in which if the difference between the two signals is within an allowable value, the higher value is the true value, and if the difference is greater than the allowable value, the true value is the one with the smaller rate of change.
給水制御装置１１からの給水要求信号は、信号
変換器１２に供給される。一方、信号変換器１２
には、給水ポンプ２の吐出圧力検出器２２からの
吐出圧力信号も供給される。信号変換器１２は、
前記給水要求信号および吐出圧力信号を２変数
（入力）とし、要求タービン回転数信号を出力す
る２変数関数発生機能を具備している。 A water supply request signal from the water supply control device 11 is supplied to a signal converter 12 . On the other hand, the signal converter 12
A discharge pressure signal from the discharge pressure detector 22 of the water supply pump 2 is also supplied to the water supply pump 2 . The signal converter 12 is
The water supply request signal and the discharge pressure signal are used as two variables (inputs), and a two-variable function generation function is provided to output a requested turbine rotation speed signal.
第６図は、給水流量WQ（横軸）と吐出圧力Hp
（縦軸）及びタービン速度（定格に対する％）の
関係を示すポンプ特性図である。図より明らかな
ように、要求給水流量と吐出圧力により、タービ
ン速度、したがつて、給水ポンプの回転数が一意
的に求められることがわかる。 Figure 6 shows the water supply flow rate WQ (horizontal axis) and discharge pressure Hp.
(vertical axis) and turbine speed (% of rated). As is clear from the figure, it can be seen that the turbine speed, and therefore the rotation speed of the water pump, can be uniquely determined by the required feed water flow rate and discharge pressure.
すなわち、タービン速度をＮ（％）とし、吐出
圧力をHp、給水流量をWQとすると、第６図の
特性から分るように、タービン速度Ｎは、
Ｎ＝ｇ（Hp、WQ）
の関数で表わされ、前記信号変換器１２が有する
２変数関数発生機能により求めることが出来る。 In other words, if the turbine speed is N (%), the discharge pressure is Hp, and the water supply flow rate is WQ, then as can be seen from the characteristics in Figure 6, the turbine speed N is a function of N = g (Hp, WQ). can be obtained using the two-variable function generation function of the signal converter 12.
しかし、従来は、ポンプの吐出圧力信号を使用
せず、要求給水流量WQのみからタービン速度を
求めていた。すなわち、従来は給水ポンプ２の吐
出圧力を一定とした運転状態を仮定していた。 However, conventionally, the turbine speed was determined only from the required water supply flow rate WQ without using the pump discharge pressure signal. That is, conventionally, an operating state in which the discharge pressure of the water supply pump 2 is kept constant has been assumed.
従つて、特に低流量域では、給水流量対速度の
ゲインが高いため、吐出圧力の変化を無視する
と、対応するタービン速度を正確に求めることが
出来ず、制御系が不安定となり運転が出来ないと
いう問題があつた。 Therefore, especially in the low flow area, the gain in feedwater flow rate vs. speed is high, so if changes in discharge pressure are ignored, the corresponding turbine speed cannot be determined accurately, making the control system unstable and operation impossible. There was a problem.
本発明は、前述の不安定要因をとり除き、低流
量域でも運転可能とするため、吐出圧力信号を追
加して正確なタービン速度および給水ポンプの回
転数を得ることを可能としたものである。 The present invention eliminates the above-mentioned instability factors and enables operation even in a low flow rate range by adding a discharge pressure signal to obtain accurate turbine speed and water pump rotation speed. .
偏差演算器１３では、速度信号と、真値選択器
１０からの速度真値とを比較して速度偏差を求め
る。速度偏差信号は、PI演算器１４で演算され、
その結果、タービン駆動蒸気流量Ｑが得られる。 The deviation calculator 13 compares the speed signal with the true speed value from the true value selector 10 to obtain a speed deviation. The speed deviation signal is calculated by the PI calculator 14,
As a result, a turbine drive steam flow rate Q is obtained.
駆動蒸気流量Ｑは、圧力補正器１５により油圧
サーボ位置Ｘに変換された後、増幅器１６でパワ
ー増幅される。前記信号Ｘは、さらに電気油圧変
換器１７で油圧に変換され、油圧サーボ１８を駆
動する。油圧サーボ１８の位置に対応して、リン
ク機構１９により高圧加減弁３および低圧加減弁
４が開閉される。 The driving steam flow rate Q is converted into a hydraulic servo position X by a pressure compensator 15, and then power amplified by an amplifier 16. The signal X is further converted into oil pressure by an electro-hydraulic converter 17 and drives a hydraulic servo 18. Corresponding to the position of the hydraulic servo 18, the link mechanism 19 opens and closes the high pressure regulating valve 3 and the low pressure regulating valve 4.
一方、高圧加減弁３の入口の高圧蒸気圧力を検
出器２０で、また低圧加減弁４の入口の低圧蒸気
圧力を検出器２１でそれぞれ検出し、その信号を
圧力補正器１５へ補正信号として送る。 On the other hand, the high pressure steam pressure at the inlet of the high pressure regulating valve 3 is detected by the detector 20, and the low pressure steam pressure at the inlet of the low pressure regulating valve 4 is detected by the detector 21, and the signals are sent to the pressure compensator 15 as a correction signal. .
タービン１は、通常運転時は主タービンの抽気
からの低圧蒸気を駆動蒸気とし、低圧蒸気圧力が
低下し、その結果駆動力が不足した場合にのみ、
主タービン入口の主蒸気からの高圧蒸気が加えら
れる。 During normal operation, the turbine 1 uses low-pressure steam from the extracted air of the main turbine as driving steam, and only when the low-pressure steam pressure decreases and the driving force is insufficient as a result,
High pressure steam from the main steam at the main turbine inlet is added.
高圧加減弁３および低圧加減弁４は、油圧サー
ボ１８の位置Ｘに対応して、リンク機構１９を介
して開閉される。リンク機構１９は、低圧加減弁
４が全開となつた後にはじめて高圧加減弁３が開
となる構造とされている。 The high pressure regulating valve 3 and the low pressure regulating valve 4 are opened and closed via a link mechanism 19 in accordance with the position X of the hydraulic servo 18. The link mechanism 19 is structured such that the high pressure regulating valve 3 opens only after the low pressure regulating valve 4 is fully opened.
油圧サーボ１８の位置Ｘに対する、タービン１
への流入蒸気流量Ｑの関係を第２図に示す。 Turbine 1 for position X of hydraulic servo 18
The relationship between the inflow steam flow rate Q and the flow rate Q is shown in Fig. 2.
油圧サーボ１８の位置Ｘが０から大きくなるに
つれて低圧加減弁４が開き、蒸気流量Ｑは増加す
る。Xaの位置で、低圧加減弁は全開となり、定
格の低圧蒸気圧力での最大流量Qa0が得られる。 As the position X of the hydraulic servo 18 increases from 0, the low pressure regulating valve 4 opens and the steam flow rate Q increases. At position X a , the low pressure regulating valve is fully open and the maximum flow rate Q a0 at the rated low steam pressure is obtained.
Xaを超えて、油圧サーボの位置Ｘが増加する
と高圧加減弁３が開き、蒸気流量はさらに増加す
る。Xbの位置で高圧加減弁も全開となり、定格
の高圧、低圧蒸気圧力での最大流量Qb0が得られ
る。 When the position X of the hydraulic servo increases beyond X a , the high pressure regulating valve 3 opens and the steam flow rate further increases. At position X b , the high pressure regulating valve is also fully opened, and the maximum flow rate Q b0 at the rated high pressure and low steam pressure is obtained.
なお、第２図は、高圧低圧のそれぞれの入口蒸
気圧力が定格の場合の関係であり、圧力が変動し
た場合には、明らかなように、圧力に比例して蒸
気流量は変化する。 Note that FIG. 2 shows the relationship when the inlet steam pressures of high pressure and low pressure are rated, and as is clear, when the pressure fluctuates, the steam flow rate changes in proportion to the pressure.
このために、本実施例においては、圧力補正器
１５を付設している。圧力補正器１５の特性を、
第３図に示す。第３図では、PI演算器１４の出
力として得られるタービン駆動蒸気流量Ｑを横軸
にとり、また縦軸には油圧サーボ位置Ｘをとる。 For this reason, in this embodiment, a pressure compensator 15 is provided. The characteristics of the pressure compensator 15 are
It is shown in Figure 3. In FIG. 3, the horizontal axis represents the turbine drive steam flow rate Q obtained as the output of the PI calculator 14, and the hydraulic servo position X is represented on the vertical axis.
油圧サーボ位置が０≦Ｘ≦Xaならば、低圧加
減弁の制御範囲であり、低圧蒸気圧力PLが定格
PLOの場合には、Ｘ＝fPL（Ｑ）と表わすことが出
来る。圧力が変動した場合の油圧サーボ位置は
Ｘ＝fPL（Ｑ）・PLO／PL
で得られ、蒸気圧力が小さくなるとサーボ位置を
増加して蒸気流量を確保できる。 If the hydraulic servo position is 0≦X≦X a , it is within the control range of the low pressure regulating valve, and the low pressure steam pressure P L is the rated value.
In the case of P LO , it can be expressed as X=f PL (Q). The hydraulic servo position when the pressure fluctuates is obtained by X= fPL (Q)・PLO / PL , and when the steam pressure decreases, the servo position can be increased to ensure the steam flow rate.
サーボ弁位置が低圧加減弁４の全開位置Xaを
超える流量になると、高圧加減弁３が開く。油圧
サーボ位置がXa≦Ｘ≦Xbならば、高圧加減弁の
制御範囲であり、高圧蒸気圧力PHが定格のPHOの
場合には、
Ｘ＝fPH（QH）＋Xa
となる。ここでQHは、高圧加減弁３を流れる流
量である。圧力が変動した場合には低圧の場合と
同様
Ｘ＝fPH（QH）・PHO／PH＋Xa
となる。 When the servo valve position reaches a flow rate exceeding the fully open position X a of the low pressure regulating valve 4, the high pressure regulating valve 3 opens. If the hydraulic servo position is X a ≦ X ≦ X b , it is within the control range of the high pressure regulator, and if the high pressure steam pressure P H is the rated P HO , then . Here, Q H is the flow rate flowing through the high pressure regulating valve 3. When the pressure fluctuates, as in the case of low pressure, X=f PH (Q H )・P HO /P H +X a .
以上の圧力補正器１５の特性の演算フローを第
４図に示す。ここでQa0は、定格の低圧蒸気圧力
における低圧加減弁全開時の蒸気流量である。 The calculation flow of the characteristics of the pressure corrector 15 described above is shown in FIG. Here, Q a0 is the steam flow rate when the low pressure regulating valve is fully open at the rated low steam pressure.
次に、圧力補正器１５の前述の特性を、演算器
（ハードウエア）で実現した一例を第５図に示す。 Next, FIG. 5 shows an example in which the above-mentioned characteristics of the pressure corrector 15 are realized by a computing unit (hardware).
PI演算器１４からの蒸気流量Ｑの信号は、関
数発生器１５０により、低圧蒸気圧力PLが定格
PLOの場合の油圧サーボ位置Ｘ＝fPL（Ｑ）に変換
される。得られた油圧サーボ位置Ｘは、さらに除
算器１５１において、低圧蒸気圧力PLで補正さ
れた位置
Ｘ＝fPL（Ｑ）・PLO／PL
に変換される。補正位置Ｘは、制限器１５２によ
つて、Ｘ≧XaのときはXaに制限される。また、
Ｘ≧Xaのときは、比較器１５３の出力により、
接点１５９が閉成される。 The signal of the steam flow rate Q from the PI calculator 14 is determined by the function generator 150 to determine the rated low pressure steam pressure P L.
The hydraulic servo position in the case of P LO is converted to X = f PL (Q). The obtained hydraulic servo position X is further converted by a divider 151 into a position X=f PL (Q)·P LO / PL corrected by the low pressure steam pressure PL . The correction position X is limited by the limiter 152 to X a when X≧X a . Also,
When X≧X a , the output of the comparator 153 gives
Contact 159 is closed.
一方、信号設定器１５４では、定格の低圧蒸気
圧力PLOにおける低圧加減弁全開時の蒸気流量Qa0
を設定する。したがつて、乗算器１５５の出力と
して、低圧蒸気圧力PLによつて補正された低圧
加減弁全開時の蒸気流量Qa0・PL／PLOが得られる。 On the other hand, the signal setting device 154 determines the steam flow rate Q a0 when the low pressure regulating valve is fully open at the rated low pressure steam pressure P LO .
Set. Therefore, the output of the multiplier 155 is the steam flow rate Q a0 ·P L /P LO when the low pressure regulating valve is fully open, which is corrected by the low pressure steam pressure P L .
減算器１５６では、（Ｑ−Qa0・PL／PLOを演算し、
高圧加減弁で負担する必要流量QHを求める。関
数発生器１５７および除算器１５８では、低圧側
と同様にして、高圧蒸気圧力PHで補正された油
圧サーボ位置fPH（QH）・PHO／PHが得られる。 The subtracter 156 calculates (Q-Q a0・P L /P LO) to find the required flow rate Q H to be borne by the high pressure regulating valve.The function generator 157 and the divider 158 calculate Hydraulic servo position f PH (Q H )・P HO /P H corrected by high pressure steam pressure P H can be obtained.
接点１５９は、前述のように、高圧蒸気加減弁
３を開くべき必要流量のときは閉じ、低圧加減弁
のみで流量が確保可能なときは開かれる。加算器
１６０では、低圧加減弁と高圧加減弁の位置信号
を加え、最終的な油圧サーボ位置Ｘが得られる。 As described above, the contact 159 is closed when the flow rate is required to open the high pressure steam control valve 3, and is opened when the flow rate can be secured only by the low pressure control valve. The adder 160 adds the position signals of the low pressure regulating valve and the high pressure regulating valve to obtain the final hydraulic servo position X.
以上のように本実施例による各種圧力補正を追
加することにより、吐出圧力の変動時の正確な所
要回転数を求めると同時に、蒸気圧力変動の結果
が回転数に表われる前に、サーボ位置を補正する
ことにより、圧力変動の影響を受けない安定した
運転が実現される。 As described above, by adding various pressure corrections according to this embodiment, the accurate required rotation speed when the discharge pressure fluctuates can be determined, and at the same time, the servo position can be determined before the result of steam pressure fluctuation appears in the rotation speed. By making the correction, stable operation that is not affected by pressure fluctuations is realized.
特に、送電線事故などにより、発電所の主ター
ビンの負荷を急激に絞つて所内単独運転を行う場
合、低流量域での所要回転数を求めると共に低圧
蒸気圧力の低下を検出し、速かに高圧加減弁を開
くことが出来る。これによつて、所内単独運転に
見合う給水量を確保することが可能となる。 In particular, when the load on the main turbine of a power plant is suddenly reduced and the plant is operated independently due to an accident on a power transmission line, etc., the required rotation speed in the low flow area is determined, and a drop in low pressure steam pressure is detected, and the High pressure regulating valve can be opened. This makes it possible to secure a water supply amount commensurate with in-house independent operation.
なお、前述の油圧サーボ１８の代りに通常のサ
ーボ機構や弁駆動装置が用い得ることは当然であ
る。 It goes without saying that a normal servo mechanism or valve drive device can be used instead of the above-mentioned hydraulic servo 18.
また、前述のタービン流入蒸気圧力による弁開
度目標値の補正は、必ずしも必要ではなく、これ
を省略することも可能である。 Further, the above-mentioned correction of the valve opening target value based on the turbine inflow steam pressure is not necessarily necessary and can be omitted.
第７図は、本発明の他の実施例のブロツク図で
ある。図において、第１図と同一の符号は同一ま
たは同等部分をあらわしている。 FIG. 7 is a block diagram of another embodiment of the invention. In the figure, the same reference numerals as in FIG. 1 represent the same or equivalent parts.
１００はモータ、１０１は前記モータ１００と
給水ポンプ２との間に介在された流体継手、１４
Ａは給水ポンプの回転数偏差（偏差演算器１３の
出力）を供給されて、前記流体継手１０１の結合
度制御信号を出力する演算器である。 100 is a motor; 101 is a fluid coupling interposed between the motor 100 and the water supply pump 2; 14;
A is a computing unit that is supplied with the rotational speed deviation of the water supply pump (output of the deviation computing unit 13) and outputs a coupling degree control signal for the fluid coupling 101.
動作時には、まず、前述と同様にして、給水制
御装置１１からの給水要求信号および吐出圧力検
出器２２からの吐出圧力に基づいて、信号変換器
１２から給水ポンプの目標回転数信号が出力され
る。 During operation, first, in the same manner as described above, a target rotation speed signal of the water supply pump is output from the signal converter 12 based on the water supply request signal from the water supply control device 11 and the discharge pressure from the discharge pressure detector 22. .
この目標回転数信号は、偏差演算器１３におい
て、給水ポンプの実測回転数と比較され、偏差が
演算器１４Ａに入力される。演算器１４Ａでは、
前記偏差に基づいて、流体継手１０１の結合度操
作量が演算され、出力される。 This target rotational speed signal is compared with the actually measured rotational speed of the water supply pump in the deviation calculator 13, and the deviation is input to the calculator 14A. In the computing unit 14A,
Based on the deviation, the degree of coupling operation amount of the fluid coupling 101 is calculated and output.
前記結合度操作量に基づいて、流体継手１０１
の結合度が操作され、給水ポンプ２の回転数が目
標値に合致させられる。 Based on the degree of coupling operation amount, the fluid coupling 101
The degree of coupling is manipulated, and the rotational speed of the water supply pump 2 is made to match the target value.
この実施例においても、給水ポンプの回転数は
全流量域にわたつて正確に制御される。なお、こ
の実施例では、回転数のフイードバツクルーブは
省略することもできる。 Also in this embodiment, the rotational speed of the water pump is accurately controlled over the entire flow range. Note that in this embodiment, the rotational speed feedback loop can be omitted.
以上の説明から明らかなように、本発明によれ
ば、給水運転時の、特に低給水流量域で高ゲイン
となる給水対速度特性を改善できるので、吐出圧
力が変動した時でも給水ポンプの安定した運転が
実現される。 As is clear from the above explanation, according to the present invention, it is possible to improve the water supply vs. speed characteristic during water supply operation, which has a high gain, especially in the low water flow rate region, so that the water supply pump can be stabilized even when the discharge pressure fluctuates. This allows for efficient driving.
第１図は本発明の一実施例を示す系統図、第２
図は油圧サーボ位置とタービン流入蒸気量の関係
を示した図、第３図は本発明の前記実施例におけ
る圧力補正特性を示した図、第４図は本発明の前
記実施例における圧力補正の演算フローを示した
図、第５図は圧力補正器のブロツク図、第６図は
給水ポンプの給水流量−吐出圧力特性を示した
図、第７図は本発明の他の実施例を示すブロツク
図である。
１……蒸気タービン、２……給水ポンプ、３…
…高圧蒸気加減弁、４……低圧蒸気加減弁、６，
７……速度検出器、１１……給水制御装置、１３
……偏差演算器、１４……PI演算器、１５……
圧力補正器、１７……電気油圧変換器、１８……
油圧サーボ、１９……リンク機構、２０……高圧
蒸気圧力検出器、２１……低圧蒸気圧力検出器、
２２……吐出圧力検出器。
Figure 1 is a system diagram showing one embodiment of the present invention, Figure 2 is a system diagram showing an embodiment of the present invention.
The figure shows the relationship between the hydraulic servo position and the amount of steam flowing into the turbine, FIG. 3 shows the pressure correction characteristics in the embodiment of the present invention, and FIG. 4 shows the pressure correction characteristics in the embodiment of the invention. FIG. 5 is a block diagram of the pressure compensator; FIG. 6 is a diagram showing the water supply flow rate vs. discharge pressure characteristic of the water supply pump; FIG. 7 is a block diagram showing another embodiment of the present invention. It is a diagram. 1...Steam turbine, 2...Water pump, 3...
...High pressure steam control valve, 4...Low pressure steam control valve, 6,
7...Speed detector, 11...Water supply control device, 13
... Deviation calculator, 14... PI calculator, 15...
Pressure compensator, 17... Electro-hydraulic converter, 18...
Hydraulic servo, 19... link mechanism, 20... high pressure steam pressure detector, 21... low pressure steam pressure detector,
22...Discharge pressure detector.
Claims (1)
速度の間の給水ポンプ特性関数を有する２変数関
数発生器に、給水要求値および給水ポンプの吐出
圧力信号を供給して前記給水ポンプの速度指令を
求め、前記速度指令に基づいて前記給水ポンプ駆
動手段の回転速度を調整制御することを特徴とす
る給水ポンプの回転数制御方式。 ２ 給水ポンプ駆動手段が蒸気タービンであり、
蒸気タービンへの供給蒸気量が前記速度指令に基
づいて調整制御されることを特徴とする前記特許
請求の範囲第１項記載の給水ポンプの回転数制御
方式。 ３ 給水ポンプ駆動手段が、モータおよびこれに
連結された流体継手であり、前記流体継手の結合
度が前記速度指令に基づいて調整制御されること
を特徴とする前記特許請求の範囲第１項記載の給
水ポンプの回転数制御方式。 ４ 給水ポンプの吐出圧力、給水流量および回転
速度の間の給水ポンプ特性関数を有する２変数関
数発生器に、給水要求値および給水ポンプの吐出
圧力信号を供給して前記給水ポンプの速度指令を
求め、前記速度指令に基づいて前記給水ポンプを
駆動するタービンへの供給蒸気量を調整制御する
ようにした給水ポンプの回転数制御方式におい
て、 前記速度指令からの速度実測値の偏差に基づい
てタービン駆動蒸気流量を求め、さらに前記ター
ビン駆動蒸気流量をタービン流入蒸気圧力で補正
して蒸気加減弁位置信号を求め、この蒸気加減弁
位置信号に基づいて蒸気加減弁の開度を制御する
ことを特徴とする給水ポンプの回転数制御方式。[Scope of Claims] 1. Supplying the water supply request value and the water supply pump discharge pressure signal to a two-variable function generator having a water supply pump characteristic function between the discharge pressure, the water supply flow rate, and the rotational speed of the water supply pump to generate the water supply. A water supply pump rotation speed control method, characterized in that a pump speed command is obtained, and the rotation speed of the water supply pump driving means is adjusted and controlled based on the speed command. 2. The water supply pump driving means is a steam turbine,
2. The water pump rotation speed control system according to claim 1, wherein the amount of steam supplied to the steam turbine is adjusted and controlled based on the speed command. 3. The water supply pump driving means is a motor and a fluid coupling connected to the motor, and the degree of coupling of the fluid coupling is adjusted and controlled based on the speed command. Water supply pump rotation speed control method. 4 Supplying the water supply request value and the water supply pump discharge pressure signal to a two-variable function generator having a water supply pump characteristic function between the water supply pump discharge pressure, the water supply flow rate, and the rotational speed to determine the speed command of the water supply pump. , in a feed water pump rotation speed control system that adjusts and controls the amount of steam supplied to a turbine that drives the feed water pump based on the speed command, the turbine drive being controlled based on the deviation of the actual speed value from the speed command. A steam flow rate is determined, and the steam control valve position signal is determined by correcting the turbine driving steam flow rate with turbine inflow steam pressure, and the opening degree of the steam control valve is controlled based on this steam control valve position signal. A water pump rotation speed control method.
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP1325182A JPS58133504A (en) | 1982-02-01 | 1982-02-01 | Control system of number of revolution of feed pump |
Applications Claiming Priority (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP1325182A JPS58133504A (en) | 1982-02-01 | 1982-02-01 | Control system of number of revolution of feed pump |
Publications (2)
| Publication Number | Publication Date |
|---|---|
| JPS58133504A JPS58133504A (en) | 1983-08-09 |
| JPH0339202B2 true JPH0339202B2 (en) | 1991-06-13 |
Family
ID=11827985
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| JP1325182A Granted JPS58133504A (en) | 1982-02-01 | 1982-02-01 | Control system of number of revolution of feed pump |
Country Status (1)
| Country | Link |
|---|---|
| JP (1) | JPS58133504A (en) |
Families Citing this family (2)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPS59150205A (en) * | 1983-02-16 | 1984-08-28 | 株式会社東芝 | Controller for feedpump |
| JPS61128508U (en) * | 1985-01-28 | 1986-08-12 |
Citations (5)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPS53136105A (en) * | 1977-05-02 | 1978-11-28 | Toshiba Corp | Feed water pump actuating turbine controller |
| JPS5490404A (en) * | 1977-12-27 | 1979-07-18 | Toshiba Corp | Feed water flow rate controller |
| JPS567904A (en) * | 1979-06-27 | 1981-01-27 | Tokyo Shibaura Electric Co | Method of controlling water pump in running under only station load |
| JPS5647606A (en) * | 1979-09-26 | 1981-04-30 | Mitsubishi Heavy Ind Ltd | Steam turbine plant |
| JPS5712202A (en) * | 1980-06-24 | 1982-01-22 | Tokyo Shibaura Electric Co | Turbine controller |
-
1982
- 1982-02-01 JP JP1325182A patent/JPS58133504A/en active Granted
Patent Citations (5)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPS53136105A (en) * | 1977-05-02 | 1978-11-28 | Toshiba Corp | Feed water pump actuating turbine controller |
| JPS5490404A (en) * | 1977-12-27 | 1979-07-18 | Toshiba Corp | Feed water flow rate controller |
| JPS567904A (en) * | 1979-06-27 | 1981-01-27 | Tokyo Shibaura Electric Co | Method of controlling water pump in running under only station load |
| JPS5647606A (en) * | 1979-09-26 | 1981-04-30 | Mitsubishi Heavy Ind Ltd | Steam turbine plant |
| JPS5712202A (en) * | 1980-06-24 | 1982-01-22 | Tokyo Shibaura Electric Co | Turbine controller |
Also Published As
| Publication number | Publication date |
|---|---|
| JPS58133504A (en) | 1983-08-09 |
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