【摘要】：直驅(qū)式電液伺服系統(tǒng)是一種新型的電液伺服系統(tǒng)。在新型的直驅(qū)式電液伺服系統(tǒng)中,電機(jī)即作為系統(tǒng)的能量元件驅(qū)動(dòng)雙向定量泵轉(zhuǎn)動(dòng)帶動(dòng)負(fù)載運(yùn)動(dòng),又作為系統(tǒng)的控制元件通過控制電機(jī)的轉(zhuǎn)速和旋轉(zhuǎn)方向來控制雙向定量泵的轉(zhuǎn)速和旋轉(zhuǎn)方向來控制系統(tǒng)液壓油的流速和循環(huán)方向,進(jìn)而控制負(fù)載運(yùn)動(dòng)。該種系統(tǒng)具有體積小、耗能低、噪聲低、效率高、控制靈活等諸多優(yōu)點(diǎn)在航空、航天、航海領(lǐng)域有著廣泛的應(yīng)用空間。隨著系統(tǒng)的不斷應(yīng)用,如何提高系統(tǒng)低速運(yùn)動(dòng)的性能已經(jīng)成為了電液伺服系統(tǒng)的研究的一個(gè)重要方向。在新型的直驅(qū)式電液伺服系統(tǒng)中,研究系統(tǒng)低速特性因素對(duì)系統(tǒng)的影響及針對(duì)各種影響因素所應(yīng)實(shí)施的補(bǔ)償方法就變得尤為重要。在本文中,建立了交流異步電機(jī)運(yùn)動(dòng)方程,直取式電液伺服系統(tǒng)液壓動(dòng)力機(jī)構(gòu)運(yùn)動(dòng)方程,分別得到了交流異步電機(jī)與液壓動(dòng)力機(jī)構(gòu)的傳遞函數(shù),并求得了直驅(qū)式電液伺服系統(tǒng)的傳遞函數(shù)�；赟imulink軟件平臺(tái)建立了直接轉(zhuǎn)矩控制異步電機(jī)仿真模型。同時(shí),也建立了基于AMEsim軟件平臺(tái)的液壓動(dòng)力機(jī)構(gòu)仿真模型,并通過兩部分的合并,建立了直驅(qū)式電液伺服系統(tǒng)聯(lián)合仿真模型。進(jìn)行了理想狀態(tài)下直驅(qū)式電液伺服系統(tǒng)的典型輸入仿真,得到了理想狀態(tài)下系統(tǒng)對(duì)典型輸入的響應(yīng)曲線,驗(yàn)證了系統(tǒng)的穩(wěn)定性。在本文中,分析了摩擦干擾力矩、齒輪泵容積損耗、齒輪泵機(jī)械損耗以及電機(jī)低速旋轉(zhuǎn)狀態(tài)下的轉(zhuǎn)矩脈動(dòng)等等各種影響系統(tǒng)低速性能的因素,分別建立了數(shù)學(xué)模型。選擇LuGre摩擦模型建立摩擦干擾力矩的仿真模型、建立了以齒輪泵端面間隙泄露與齒輪泵徑向間隙泄露為主的齒輪泵容積損耗仿真模型、建立了以齒輪泵齒頂端面與液體的粘性摩擦損失為主的齒輪泵機(jī)械損耗仿真模型。分別就各個(gè)因素注入到直驅(qū)式電液伺服系統(tǒng)理想狀態(tài)下仿真模型中進(jìn)行對(duì)比仿真,觀察并分析了各因素對(duì)直驅(qū)式電液伺服系統(tǒng)的影響。在本文中,通過分析各因素對(duì)直驅(qū)式電液伺服系統(tǒng)的影響,選擇了高增益PID控制器與反步積分自適應(yīng)控制器分別對(duì)摩擦力矩進(jìn)行了補(bǔ)償。建立了高增益PID控制器與反步積分自適應(yīng)控制器的數(shù)學(xué)模型,并在上述模型的基礎(chǔ)上分別建立了高增益PID控制器與反步積分自適應(yīng)控制器的仿真模型,將其分別注入到含摩擦干擾力矩的直驅(qū)式電液伺服系統(tǒng)仿真模型中,建立經(jīng)過補(bǔ)償?shù)暮Σ粮蓴_力矩的直驅(qū)式電液伺服系統(tǒng)仿真模型。通過高增益PID控制器補(bǔ)償和反步積分自適應(yīng)控制器補(bǔ)償兩種方法的對(duì)比仿真,驗(yàn)證了反步積分自適應(yīng)控制器對(duì)摩擦干擾力矩的補(bǔ)償效果更加有效,補(bǔ)償效果符合要求。針對(duì)齒輪泵容積損耗問題,針對(duì)齒輪泵容積損耗的特點(diǎn)設(shè)計(jì)了物理補(bǔ)油裝置,并針對(duì)補(bǔ)油裝置和液壓鎖閥設(shè)計(jì)了集成閥塊。完成了直驅(qū)式電液伺服系統(tǒng)低速控制的研究?jī)?nèi)容。
[Abstract]:Direct drive electro-hydraulic servo system is a new type of electro-hydraulic servo system. In a new type of direct-drive electro-hydraulic servo system, the motor is used as the energy element of the system to drive the bidirectional quantitative pump rotation and drive the load movement. As the control element of the system, the speed and direction of rotation of the bidirectional quantitative pump are controlled by controlling the speed and the direction of rotation of the motor to control the velocity and circulation direction of the hydraulic oil of the system, and then to control the movement of the load. The system has many advantages, such as small volume, low energy consumption, low noise, high efficiency, flexible control and so on. With the continuous application of the system, how to improve the performance of the system at low speed has become an important research direction of the electro-hydraulic servo system. In a new type of direct-drive electro-hydraulic servo system, it is very important to study the influence of the low speed characteristic factors on the system and the compensation method for various factors. In this paper, the equation of motion of AC asynchronous motor and the equation of motion of hydraulic power mechanism of direct electro-hydraulic servo system are established, and the transfer functions of AC asynchronous motor and hydraulic power mechanism are obtained respectively. The transfer function of direct-drive electro-hydraulic servo system is obtained. The simulation model of direct torque control asynchronous motor is established based on Simulink software platform. At the same time, the simulation model of hydraulic power mechanism based on AMEsim software platform is established, and the joint simulation model of direct-drive electro-hydraulic servo system is established by combining the two parts. The typical input simulation of direct-drive electro-hydraulic servo system in ideal state is carried out. The response curve of the system to typical input in ideal state is obtained and the stability of the system is verified. In this paper, the factors that affect the low speed performance of the system, such as friction disturbance moment, gear pump volume loss, gear pump mechanical loss and torque ripple under the condition of motor low speed rotation, are analyzed, and the mathematical models are established respectively. The LuGre friction model is selected to establish the simulation model of friction disturbance moment, and the simulation model of gear pump volume loss is established, which is mainly based on the leakage of the end clearance of gear pump and the leakage of radial clearance of gear pump. The mechanical loss simulation model of gear pump is established, which is based on the viscous friction loss between the top surface of gear pump tooth and liquid. Each factor is injected into the ideal simulation model of direct-drive electro-hydraulic servo system, and the influence of each factor on direct-drive electro-hydraulic servo system is observed and analyzed. In this paper, by analyzing the influence of various factors on the direct-drive electro-hydraulic servo system, the high gain PID controller and the backstepping integral adaptive controller are selected to compensate the friction torque respectively. The mathematical models of high gain PID controller and backstepping integral adaptive controller are established, and the simulation models of high gain PID controller and backstepping integral adaptive controller are established based on the above models. It is injected into the simulation model of direct-drive electro-hydraulic servo system with friction disturbance torque, and the simulation model of direct-drive electro-hydraulic servo system with friction disturbance moment is established. Through the comparison and simulation of high gain PID controller compensation and backstepping integral adaptive controller compensation, it is proved that the backstepping integral adaptive controller is more effective to compensate friction disturbance torque, and the compensation effect meets the requirements. Aiming at the problem of gear pump volume loss, the physical oil filling device is designed according to the characteristics of gear pump volume loss, and the integrated valve block is designed for oil filling device and hydraulic lock valve. The research content of low speed control of direct drive electro-hydraulic servo system is completed.
【學(xué)位授予單位】：哈爾濱工程大學(xué)
【學(xué)位級(jí)別】：碩士
【學(xué)位授予年份】：2014
【分類號(hào)】：TM921.541

【參考文獻(xiàn)】

相關(guān)期刊論文 前10條

1 楊存智;劉文藝;;電液比例伺服控制容積調(diào)速系統(tǒng)仿真研究[J];機(jī)床與液壓;2012年13期

2 周金柱;段寶巖;黃進(jìn);;LuGre摩擦模型對(duì)伺服系統(tǒng)的影響與補(bǔ)償[J];控制理論與應(yīng)用;2008年06期

3 陳英;荊寶德;王義強(qiáng);;外嚙合齒輪泵內(nèi)泄漏理論模型的建立及參數(shù)優(yōu)化[J];機(jī)床與液壓;2007年10期

4 陳英;荊寶德;王義強(qiáng);;外嚙合齒輪泵的間隙優(yōu)化[J];機(jī)床與液壓;2007年09期

5 劉海榮;劉金琨;;Lugre摩擦模型的模糊神經(jīng)網(wǎng)絡(luò)辨識(shí)仿真研究[J];計(jì)算機(jī)仿真;2007年01期

6 徐繩武;從節(jié)能看液壓傳動(dòng)控制系統(tǒng)發(fā)展的三個(gè)階段[J];液壓氣動(dòng)與密封;2005年05期

7 姜繼海,涂婉麗,曹健;直驅(qū)式容積控制電液伺服系統(tǒng)動(dòng)態(tài)性能研究[J];液壓與氣動(dòng);2005年08期

8 姜繼海,蘇文海,張洪波,劉慶和;直驅(qū)式容積控制電液伺服系統(tǒng)及其在船舶舵機(jī)上的應(yīng)用[J];中國造船;2004年04期

9 吳秋生;交流傳動(dòng)的應(yīng)用與展望[J];自動(dòng)化博覽;2004年02期

10 姜繼海,蘇文海,劉慶和;直驅(qū)式容積控制電液伺服系統(tǒng)[J];軍民兩用技術(shù)與產(chǎn)品;2003年09期

相關(guān)博士學(xué)位論文 前3條

1 鄭洪波;伺服直驅(qū)泵控液壓系統(tǒng)及其節(jié)能機(jī)理研究[D];廣東工業(yè)大學(xué);2012年

2 蘇文海;直驅(qū)式電液伺服轉(zhuǎn)葉舵機(jī)關(guān)鍵技術(shù)及其控制系統(tǒng)研究[D];哈爾濱工業(yè)大學(xué);2009年

3 李強(qiáng);并聯(lián)電液伺服六自由度平臺(tái)系統(tǒng)低速運(yùn)動(dòng)研究[D];浙江大學(xué);2008年

相關(guān)碩士學(xué)位論文 前7條

1 韓洪祥;船舶直驅(qū)式容積控制液壓舵機(jī)系統(tǒng)的分析與設(shè)計(jì)[D];哈爾濱工程大學(xué);2010年

2 趙朝星;異步電動(dòng)機(jī)直接轉(zhuǎn)矩控制系統(tǒng)及性能改善研究[D];中南大學(xué);2008年

3 張丹;含摩擦環(huán)節(jié)伺服系統(tǒng)的補(bǔ)償控制[D];西安電子科技大學(xué);2008年

4 劉強(qiáng);混合動(dòng)力汽車感應(yīng)電動(dòng)機(jī)模糊直接轉(zhuǎn)矩控制系統(tǒng)的研究[D];沈陽工業(yè)大學(xué);2008年

5 馬艷玲;含齒隙環(huán)節(jié)伺服系統(tǒng)的補(bǔ)償控制[D];西安電子科技大學(xué);2008年

6 楊興武;異步電動(dòng)機(jī)直接轉(zhuǎn)矩調(diào)速系統(tǒng)的設(shè)計(jì)與仿真研究[D];貴州大學(xué);2007年

7 劉洪玉;轉(zhuǎn)臺(tái)伺服系統(tǒng)低速性能分析與摩擦補(bǔ)償研究[D];哈爾濱工業(yè)大學(xué);2006年

，

本文編號(hào)：2411932