中英翻譯離心式和往復(fù)式壓縮機(jī)的工作效率特性

1、離心式和往復(fù)式壓縮機(jī)的工作效率特性Rainer Kurz , Bern hard Win kelma nn , and Saeid Mokhatab往復(fù)式壓縮機(jī)和離心式壓縮機(jī)具有不同的工作特性，而且關(guān)于效率的定義也不同。本文提供了一個公平的比較準(zhǔn)則，得到了對于兩種類型機(jī)器普遍適用的效率定義。 這個比較基于用戶最感興趣的要求提出的。 此外，對于管道的工作環(huán)境影響和在 不同負(fù)載水平的影響給出了評估。乍一看，計算任何類型的壓縮效率看似是很簡單的：比較理想壓縮過程和實(shí)際壓縮過程的工作效率。難點(diǎn)在于正確定義適當(dāng)?shù)南到y(tǒng)邊界， 包括與之相關(guān)的壓縮過 程的損失。除非這些邊界是恰好定義的，否則離心式和往復(fù)式壓縮

2、機(jī)的比較就變 得有缺陷了。我們也需要承認(rèn)，效率的定義，甚至是在評估公平的情況下，仍不能完全回應(yīng)操 作員的主要關(guān)心問題：壓縮過程所需的驅(qū)動力量是什么？要做到這一點(diǎn)，就需要討論在壓縮過程中的機(jī)械損失。 隨著時間的推移效率趨勢也應(yīng)被考慮， 如非設(shè) 計條件，它們是由專業(yè)的流水線規(guī)定，或者是受壓縮機(jī)的工作時間和自身退化的 影響。管道使用的壓縮設(shè)備涉及到往復(fù)式和離心式壓縮機(jī)。離心式壓縮機(jī)用燃?xì)廨?機(jī)或者是電動馬達(dá)來驅(qū)動。所用的燃?xì)廨啓C(jī)，總的來說，是兩軸發(fā)動機(jī)，電動馬 達(dá)使用的是變速馬達(dá)或者變速齒輪箱。往復(fù)壓縮機(jī)是低速整體單位或者是可分的“高速”單位，其中低速整體單位是燃?xì)獍l(fā)動機(jī)和壓縮機(jī)在一個曲柄套管內(nèi)。后

3、者單位的運(yùn)行在750-1,200rpm范圍內(nèi)(1,800rpm是更小的單位)并且通常都是 由電動馬達(dá)或者四沖程燃?xì)獍l(fā)動機(jī)來驅(qū)動。效率要確定任何壓縮過程的等熵效率，就要基于測量的壓縮機(jī)吸入和排出的總焓 變?yōu)椋?S)，于是等熵效率，總壓力(P)，溫度(T)和熵(h)sh(p,s) h(P,T) suctsuctdischsuctsh(p,T) h(p,T)(Eq.1)suctdischsuctdisch并且加上測量的穩(wěn)態(tài)質(zhì)量流 m，吸收軸功率為：mh(p,T) h(p,p T) I suctdischdischsuct(Eq.2)m?？紤]機(jī)械效率m理論(熵)功耗(這是絕熱系統(tǒng)可能出現(xiàn)的最低功耗)如

4、下：P mh(p,s) h(p,T) (Eq.3)sucttheordischsuctsuc 環(huán)境的熱交換通?？梢院雎??！胺€(wěn)態(tài)”流入和流出離心 式壓縮機(jī)的流量可以視為。系統(tǒng)邊界要包需要確定的是，系統(tǒng)邊界的效率計算通 常是用吸入和排出的噴嘴。機(jī)械效率尤其是從平衡活塞式或分裂墻滲漏的循環(huán)路 徑。含所有內(nèi)部泄露途徑，。99%和98%，在描述軸承和密封件的摩擦損失以及風(fēng)阻損失時可以達(dá)到m給出的，鑒于吸力緩沖器Eq.3對于往復(fù)式壓縮機(jī)，理論的氣體馬力也是由從臨 往復(fù)壓縮機(jī)就其性質(zhì)而言，上游和排力緩沖器下游的吸氣和排氣壓力脈動。，以及近單位需要多方面的系統(tǒng)來控制脈動和提供隔離(包括往復(fù)式和離心式)對于 任

5、何一個低速或高速單位的可以自然存在的來自管線的管流量和面積管道。歧管系統(tǒng)設(shè)計，使用了卷相結(jié)合，管道長度和壓力降元素來創(chuàng)造脈動(聲波)濾波器。這些歧管系統(tǒng)（過濾器）引起壓力下降，因此必須在效率計算時考慮到。就像離 心壓從吸氣壓力扣除的額外壓力不得不包含進(jìn)殘余脈動的影響。潛在的，縮機(jī)一樣，傳熱就經(jīng)常被忽視。對于可分機(jī)機(jī)械效率一般使用 95%對于積分的機(jī)器， 機(jī)械效率一般取為8-15%往復(fù)式發(fā)動機(jī)機(jī)械損失在這些數(shù)字似乎有些樂觀，一系列數(shù)字顯示，97%。，K.（參考1往復(fù)壓縮機(jī)招致號碼：庫爾茲，R.之間，往復(fù) 壓縮機(jī)的在6-12% ）。光布倫，2007工作環(huán)境LdLu上游和在這樣的情況下， 當(dāng)壓縮機(jī)在

6、一個系統(tǒng)中運(yùn)行時，管道長度在 pe下游的終止壓力均被視為常量， 下游，以及管道pu上游的初始壓力和管道）。管道系統(tǒng)中我們有一個壓縮機(jī)運(yùn) 行的簡單模型（圖1年）。M.由羅穆斯基，2006圖1:管道段的概念模型（文獻(xiàn) 2:庫爾茲.R，p,在壓縮機(jī)對于給定的，標(biāo)準(zhǔn)管線定量流動能力將在吸入階段強(qiáng) 加壓力sHp）關(guān)系可以）流（。對于給定的管線，壓縮機(jī)站頭部（Q放電區(qū)強(qiáng)加壓力sdk1k1 1 CTH 近似表述為（Eq.4）sps2QC C 43 12pd CC是常數(shù)（對于一個給定的管道幾何）分別描述了管道兩邊的壓力和和其中34摩擦損失（文獻(xiàn)2:庫爾茲.R,M.由羅穆斯基，2006 年）。除去其他問題，這意

7、味著對于帶管道系統(tǒng)的壓縮機(jī)站， 頭部所需流量揚(yáng)程是由管 道系統(tǒng)規(guī)定的（圖2）。特別地，這一特點(diǎn)對于壓縮機(jī)需要的能力允許頭部減量， 按照規(guī)定的方式反之亦然。管道因此將不需要改變頭部的流量恒定（或壓力比）。 圖2:建立在4式上的機(jī)頭流量關(guān)系。在短暫的情況下（如包裝其間），最初的操作條件遵循恒功率分布，如頭部流量 關(guān)系如下：H s con stP m（ Eq.5） s con st H 一s Q年）2002庫爾茲，R.，S.：奧海寧3并將漸進(jìn)地達(dá)到穩(wěn)定的關(guān)系 （文獻(xiàn)該系統(tǒng)需求的必須控制壓縮機(jī)輸出與系統(tǒng)要求匹配。在上述要求的基礎(chǔ)上，管線 壓縮機(jī)提供了在操作條件特點(diǎn)是系統(tǒng)流程和系統(tǒng)頭部或壓力比的強(qiáng)烈關(guān)系

8、。具體一個重要問題就是如何使壓縮機(jī)適應(yīng)這樣變化的條件，經(jīng)驗下的大量變化，的說就是如何影響效率。這意味著壓力比的改變對機(jī)器離心壓縮機(jī)具有相當(dāng)大的平 頭部和流程特點(diǎn)。對于一個恒速運(yùn)行年）R.，20004的實(shí)際流程有重大的影響（文獻(xiàn)4:庫爾茲控制壓縮機(jī)內(nèi)的流程可以實(shí)現(xiàn)頭部或壓力比隨著流量的增加而 減少。的壓縮機(jī)，兩軸燃?xì)廨啓C(jī)和變這是控制離心壓縮機(jī)最便捷的方法。壓縮機(jī)不同的運(yùn)行速度。50%到100%速電機(jī)允許大范圍的速度變化（通常是最大速 度或更多的40%或（受但速度是間接平衡由渦輪產(chǎn)生的動力應(yīng)當(dāng)指出，被控制的值通常不是速度， 進(jìn)入燃?xì)廨啓C(jī)燃油流量控制）和壓縮機(jī)的吸收功率。年安裝 的任何離心壓縮機(jī)在管

9、線服務(wù)方面是由調(diào)速器來事實(shí)上，在過去15年長的設(shè)施和服務(wù)設(shè)施在其他管線服務(wù)有時使用驅(qū)使的，通常是兩軸燃?xì)廨啓C(jī)。的變化）和 恒速電動機(jī)。在這些裝置中，90%到100%單軸燃?xì)廨啓C(jī)（允許速度 吸節(jié)流或可 變進(jìn)氣導(dǎo)葉用來提供控制方法。圖3:典型的管線運(yùn)行點(diǎn)繪制成的典型離心壓縮機(jī)性能圖。（窒最小流量（涌）和最大流量離心壓縮機(jī)的運(yùn)行封套受最大允許 速度限制，3）。另一個限制因素可能是可用的驅(qū)動電源。息或石墻）（圖因為它 被定義為壓縮機(jī)的一種氣動穩(wěn)定性的極只有最小流量需要特別注意，調(diào)限。跨越這個限制以降低流動將導(dǎo)致壓縮機(jī)流動逆轉(zhuǎn)，這可能會損壞壓縮機(jī)。幾乎所制解調(diào)器控制系統(tǒng)通過打開一個循環(huán)閥來控制這種情況。

10、出于這個原因，當(dāng)壓縮機(jī)內(nèi)的流量趨于穩(wěn)定極有的現(xiàn)代壓縮機(jī)裝置都使用帶有控制閥的循環(huán)線，控制系統(tǒng)不斷地監(jiān)測壓縮機(jī)關(guān)系喘振線的運(yùn)行限時這種控制閥允許流量的增加。帶有開放或部分并且有必要的話自動地開關(guān)循環(huán)閥。 對于大多數(shù)應(yīng)用來說，點(diǎn)，或者是在混 亂運(yùn)行條件時的短開放循環(huán)閥的運(yùn)行模式只被用于開啟和關(guān)閉階段，暫時期。壓縮機(jī)的葉輪將在達(dá)到或接近其最大效率時被 4得到管線特點(diǎn)，假設(shè)由公式這選 出來運(yùn)行，這個最大效率是由管線強(qiáng)加在整個系列的頭部和流量條件下的。）控制的壓縮機(jī)，因為一個壓縮機(jī)的最有效點(diǎn)是由一種關(guān)可能是有一個速度（N系而連接的，這種關(guān)系需要大約（風(fēng)扇法方程）：CHQ25 QH C（Eq.6）C5N2

11、565N2C6P是（基于效率幾乎保持不變這個的為滿足上述關(guān)系的操作點(diǎn)，吸入氣壓g事實(shí)）：C335N C C C Q C CP HQ（Eq.7）776g7552C6正因為如此，這種力-速度關(guān)系允許動力渦輪運(yùn)行達(dá)到或非常接近其整個范圍的理想速度。管線中典型的運(yùn)行方案允許壓縮機(jī)和動力渦輪在大多數(shù)時間里在 最有效點(diǎn)運(yùn)行。然而，燃?xì)廨啓C(jī)的燃?xì)馍a(chǎn)商將在部分負(fù)荷運(yùn)行時丟失一些熱效 率。不同流動要求的管線運(yùn)行點(diǎn)繪制成用于壓顯示了一個典型的實(shí)際例子：3圖縮機(jī)站中的速度控制離心壓縮機(jī)性能圖。往復(fù)壓縮機(jī)將自動服從系統(tǒng)壓力比的需求， 只要沒有超出機(jī)械的限制條件（桿負(fù) 載功率）。系統(tǒng)吸排氣壓力的改變將僅能引起閥門或早

12、或晚的開啟。頭部可以自 動下降因為閥門可以降低排氣端的管線壓力和 /或吸入端更高的管線壓力。因此, 如果沒有額外的措施，流量將大致恒定一一除了容積效率將增加的變化，所以降低壓力比而增加流量?？刂频奶魬?zhàn)存在于系統(tǒng)要求的流量調(diào)整。如果沒有額外的調(diào)整，隨著壓力比的變 化，壓縮機(jī)流量的改變微乎其微。從歷史上看，通過改變激活機(jī)器的數(shù)量使管線 安裝許多小的壓縮機(jī)和調(diào)整流量。 這個容量和負(fù)荷可通過速度調(diào)諧，或者通過一 個單一單元的缸間隙中的許多小調(diào)整（加載步驟）來調(diào)諧。隨著壓縮機(jī)的發(fā)展， 控制容量的負(fù)擔(dān)轉(zhuǎn)移到獨(dú)立壓縮機(jī)上。負(fù)荷控制是壓縮機(jī)運(yùn)行的一個關(guān)鍵組成部分。從管線操作角度來看，在機(jī)組中流 量變化要符合管

13、線投出承諾，以及實(shí)施公司最佳操作（例如，線包裝，負(fù)載預(yù)期）。 從一個單元的角度來看，負(fù)荷控制包含降低單元流量（通過卸載或速度）使操作 盡可能的貼近設(shè)計扭矩限制，并在壓縮機(jī)或驅(qū)動程序沒有超載的情況下進(jìn)行。對于任何給定的機(jī)組入口和出口壓力，在任何負(fù)荷圖曲線上的關(guān)鍵限制都是桿負(fù)荷 限制和馬力/扭矩限制。瓦斯控制通常會建立在一個機(jī)組的單元上，而這個機(jī)組 運(yùn)行必須達(dá)到管線流量目標(biāo)。地方單元控制將建立負(fù)載步驟或速度要求來限制桿 負(fù)荷或達(dá)到扭矩控制。改變流量的常用方法是改變速度，改變間隙，或取消激活缸頭（保持進(jìn)口閥開啟）。 另一種方法是卸載無限步驟，從而延緩吸氣閥封閉以減少容積效率。 此外，流程 的一部分可

14、以回收或吸氣壓力可以節(jié)流從而降低質(zhì)量流量， 同時保持進(jìn)入壓縮機(jī) 的容積流量基本不間斷。壓縮機(jī)控制策略應(yīng)該能夠?qū)崿F(xiàn)自動化， 并在壓縮機(jī)運(yùn)行期間能夠簡便地調(diào)整。 特 別地，壓縮機(jī)設(shè)計修改的戰(zhàn)略需求（如：離心壓縮機(jī)重新旋轉(zhuǎn)，改變缸徑，或給 往復(fù)壓縮機(jī)添加固定間隙）在這里不被考慮。需要指出的是，對于往復(fù)式壓縮機(jī) 一個關(guān)鍵的控制要求是不超載驅(qū)動或超過機(jī)械限制。運(yùn)行 典型的穩(wěn)態(tài)管道運(yùn)行將產(chǎn)生圖 4 所示的一個有效行為。 該圖是評估沿管道穩(wěn)定運(yùn) 行特征狀態(tài)壓縮機(jī)效率的結(jié)果。大中型壓縮機(jī)都將達(dá)到 100%流量的最佳效率， 并允許超出設(shè)計流量的 10%。不同的機(jī)械效率并沒有考慮這種對比。 往復(fù)壓縮機(jī)效率在文獻(xiàn)

15、5 中被推導(dǎo)出，從增加的閥門效率測量與壓縮效率和造成 的損失脈動衰減器。 低速壓縮機(jī)的效率是可以實(shí)現(xiàn)的。 高速往復(fù)壓縮機(jī)在效率上 可能比較低。圖 4 ：以穩(wěn)態(tài)管線特性運(yùn)行為基礎(chǔ)的在不同流量率的壓縮機(jī)效率。圖 4 顯示在較低壓力比下增加的閥門損失的影響和往復(fù)機(jī)器的較低流量， 而離心 壓縮機(jī)的效率幾乎保持常量。結(jié)論 不同型號壓縮機(jī)間的效率定義和對比需要密切關(guān)注邊界條件的定義， 對于這樣的 邊界條件， 效率和受用的運(yùn)行發(fā)展趨勢同時被定義。 當(dāng)效率值用來計算功耗時機(jī) 械效率具有重要作用。如果不考慮這些定義，不同系統(tǒng)的優(yōu)缺點(diǎn)討論將變得 不 準(zhǔn)確和有誤導(dǎo)性。參考文獻(xiàn)：1. 庫爾茲.R.K.光布倫，2007

16、?！巴鶑?fù)和離心壓縮機(jī)的效率定義和負(fù)荷管理” 美國機(jī)械工程師協(xié)會 文章 GT2007-27082. 庫爾茲.R，M.由羅穆斯基，2006。“不對稱接壓縮機(jī)站閑置產(chǎn)能”。美國機(jī) 械工程師協(xié)會 文章 2006-900693. 奧海寧.S.R.庫爾茲，2002?！皟蓹C(jī)壓縮機(jī)站的系列或平行排列”。反式。美國機(jī) 械工程師協(xié)會，第 124 欄4. 庫爾茲.R，2004。“離心壓縮機(jī)性能的物理”。管道仿真利益集團(tuán)。棕櫚泉，加 利福尼亞5. 米.瓦特沙發(fā)， 2003。“天然氣壓縮服務(wù)六主線壓縮機(jī)閥門的性能和耐用性試驗”。天然氣機(jī)械會議。鹽湖城，UT原文Efficiency And Operating Chara

17、cteristics Of Centrifugal And Reciprocating CompressorsBy Rainer Kurz, Bernhard Winkelmann, and Saeid iVIokhatabReciprocating compressors and centrifugal compressors have different operating characteristics and use different eificiency definitions. This article provides guidelines for an equitable c

18、omparison, resulting in a universal efficiency definition for both types of machines. The comparison is based on the requirements in which a user is ultimately interested. Further, the impact of actual pipeline operating conditions and the impact on efficiency at different load levels is evaluated.A

19、t first glance, calculating the efficiency for any type of compression seems to be straightforward: comparing the work required of an ideal compression process with the work required of an actual compression process. The difficulty is correctly defi ning appropriate system boun daries that in elude

20、losses associated with the compressi on process. Uni ess these boun daries are appropriately defi ned, comparis ons betwee n cen trifugal and reciprocat ing compressors become flawed.We also n eed to ack no wledge that the efficie ncy defi niti ons, eve n whe n evaluated equitably, still dont comple

21、tely answer one of the operators main concerns: What is the driver power required for the compressi on process?To accomplish this, mecha ni cal losses in the compressi on systems n eed to be discussed.Trends in efficie ncy should also be con sidered over time, such as odesig n con diti ons as they a

22、re imposed by typical pipeli ne operati ons, or the impact of operati ng hours and associated degradati on on the compressors.The compressi on equipme nt used for pipeli nes invo Ives either reciprocat ing compressors or cen trifugal compressors. Cen trifugal compressors are drive n by gas turb in e

23、s, or by electricmotors. The gas turb ines used are, in gen eral,t-shaft engines and the electric motor drives use either variable speed motors, or variable speed gearboxes. Reciprocati ng compressors are either low speed in tegral un its, which separable cas in g,or crank one in compressor the and

24、engine gas the comb ine high-speed units. The latter units operate in the 750,200 rpm range (1,800 rpm for smaller un its) and are gen erally drive n by electric motors, or fou-stroke gas engin es. Efficie ncyTo determ ine the ise ntropic efficie ncy of any compressi on process based on total en tha

25、lpies (h), total pressures (p), temperatures (T)a nd en tropies (s) at suct ion and discharge of the compressor are measured, and the ise ntropic efficie ncy rA the n becomes:h(p,s) h(p,T) suctsuctdischsuctsh(p,T) h(p,T)(Eq.1)suctdischdischsuctand, with measuring the steady state mass flow m, the ab

26、sorbed shaft power is: .mh(p,T) h(p,T)p suctsuctdischdisch(Eq.2)mconsidering the mechanical efficiency rA.The theoretical (ise ntropic) power con sumpti on (which is the lowest possible power con sumptio n for an adiabatic system) follows from: .),ph)s,( Pmhp (T(Eq.3)suctsuctsuctdischthe osteady cen

27、 trifugal The flow into and out of a compressor can be con sidered as excha nge with the en vir onment is usually n egligible. System boun daries state.Heat for the efficie ncy calculati ons are usually the sucti on and discharge no zzles. It n eeds paths,i n all intern al leakage boun daries be ass

28、ured that the system env elope to The wall leakages. fiAom balanee piston or division paths particular recirculation mechanical efficiency r)A., describing the friction losses in bearings and seals, as well as win dage losses, is typically betwee n 98 and 99%.Eq. by is also give n horsepower recipro

29、cat ing For compressors, theoretical gas pulsatio n the sucti on are suct ion 3,give n the and discharge pressure upstream ofReciprocati ng pulsati on dampe ners and dow nstream of the discharge dampe ners. compressors, by their very n ature, require manifold systems to con trol pulsati ons and prov

30、ide isolation from neighboring units (both reciprocating and centrifugal), as well nature.The extensive and can be in piping flow as from pipeline meters and yard a uses high either of desig n mani fold systems for slow speed or speed un its create comb in ati on of volumes, drop pressure eleme nts

31、to pip ing len gths and pulsation (acoustic) filters.These manifold systems (filters) cause a pressure drop, and pressure efficie ncy con sidered be in calculati ons. additi onal Pote ntially, thus must of to would suct ion from deduct ions the pressure have to made in clude effects the residual pul

32、sations. Like centrifugal compressors, heat transfer is usually neglected. For 95%. as gen erally efficie ncy mach in es, For in tegral mecha ni cal is take n separable machi nes a 97% mecha ni cal efficie ncy is ofte n used. These nu mbers seem that of nu mber sources state a that fact give n optim

33、istic, somewhat to be the reciprocati ng reciprocati ng 815% betwee n in cur engines losses mecha ni cal and compressors between 612%(Ref 1: Kurz , R., K. Brun, 2007).Operat ing Con diti onsFor a situation where a compressor operates in a system with pipe of the length Lu upstream and a pipe of the

34、len gth Ld dow nstream, and further where the pressure at are pe dow nstream pipe and the end of the the beg inning of the upstream pipe pu known and con sta nt, we have a simple model of a compressor stati on operat ing in a pipeli ne system (Figure 1).M. R., 2: Kurz, a pipeli ne segme nt (Ref. Fig

35、ure 1: Con ceptual model of Lubomirsky.2006).For a given, constant flow capacity Qstd the pipeline will then impose a pressure ps at the suct ion and pd at the discharge side of the compressor. For a give n pipeli ne, the head (Hs)-flow (Q) relatio nship at the compressor statio n can be approximate

36、d by k1k1 1 H CT(Eq.4)ssp2C C Q4312pd where C3 and C4 are constants (for a given pipelinegeometry) describ ing thepressure at either ends of the pipeli ne, and the fricti on losses, respectively(Ref 2: Kurz,R., M. Lubomirsky, 2006).Among other issues, this means that for a compressor stati on with i

37、n a pipeli ne system, the head for a required flow is prescribed by the pipeline system (Figure 2). In particular, this characteristic requires the capability for the compressors to allow a reducti on in head with reduced flow, and vice versa, in a prescribed fashi on. The pipeline will therefore no

38、t require a change in flow at constant head (or pressure ratio).Figure 2: Stafi on Head-Flow relati on ship based on Eq. 4.In tran sie nt situati ons (for example duri ng line pack in g), the operati ng con diti ons follow in itially a con sta nt power distributio n, i.e. the head flow relati on shi

39、p follows: H s m constP(Eq.5) s const1s H s Qand will asymptotically approach the steady state relati on ship (Ref 3:Ohanian, S.,R.Kurz, 2002).Based on the requirements above, the compressor output must be controlled to match the system demand. This system demand is characterized by a strong relatio

40、nship between system flow and system head or pressure ratio.Given the large variations in operating conditions experienced by pipeline compressors, an important question is how to adjust the compressor to the varying conditions, and, in particular, how does this influence the efficiency.Centrinagal

41、compressors tend to have rather flat head vs. flow characteristic. Thisflow the actual a significant effect on changes means that in pressure ratio have through the machine (Ref 4:Kurz, R., 2004). For a centrifugal compressor operating at a constant speed, the head or pressure ratio is reduced with

42、increasing flow. Controlling the flow through the compressor can be accomplished by varying the controlling method of This is the preferred operating speed of the compressor motors electric turbines and variable speed shaft centrifugal compressors. Two gas of 100% (usually from 40-50% to allow for s

43、peed variations over a wide range maximum speed or more).It should be noted, that the controlled value is usually not speed, but the speed is indirectly the result of balancing the power generated by the the and the gas turbine) flow (which is controlled by the fuel into power turbine absorbed power

44、 of the compressor.pipeline in past 15 years Virtually any centrifugal compressor installed in the Older two-shaft a gas turbine. by a variable speed driver, usually service is driven installations and installations in other than pipeline service sometimes use sing-lsehaft gas turbines (which allow

45、a speed variation from about 9-0100% speed) and constant speed electric motors. In these installations, suction throttling or variable inlet guide vanes are used to Drovide means of control.centrifugal into a typical 3: Typical pipeline operating points plotted Figure compressor performance map.maxi

46、mum by the limited of a centrifugal compressor is The operating envelope or flow (choke flow (surge flow),and the maximum allowable speed, the minimum stonewall)(Figure 3). Another limiting factor may be the available driver power. an defined by attention, because it is Only the minimum flow require

47、s special aerodynamic stability limit of the compressor Crossing this limit to lower flows will cause a flow reversal in the compressor, which can damage the compressor. Modem this situation by automatically opening a recycle valve. For control systems prevent with line installations use a recycle t

48、his reason, virtually all modern compressor control valve that allows the increase of the flow through the compressor if it comes near the stability limit. The control systems constantly monitor the operating point of the compressor in relation to its surge line,and automatically open or close the r

49、ecycle or open, mode with an if necessary. For most applications, the operating valve brief shutdown, or for is only used for start-up and partially open recycle valve periods during upset operating conditions.Assuming the pipeline characteristic derived in Eq. 4, the compressor impellers will effic

50、iency for the entire range of head and or near its best be selected to operate at flow con diti ons imposed by the pipeli ne. This is possible with a speed (N) con trolled a by a compressor are conn ected of compressor, because the best efficie ncy points relationship that requires approximately (fa

51、n law equation):CHQ25 Q HC(Eq.6)C 5N 2565N2C6Foroperati ng points that meet the above relati on ship, the absorbed gas power Pg is (due to the fact that the efficie ncy stays approximately con sta nt): C335 C Q C Q C C NP C H(Eq.7) 75767g52C6 As it is, thispower-speed relati on ship allows the power

52、 turb ine to operate at, or very close to its optimum speed for the entire range.The typical operating scenarios in pipeli nes therefore allow the compressor and the power turb ine to operate at its best eflicie ncy for most of the time. The gas producer of the gas turb ine will, however, lose some

53、thermal efficie ncy whe n operated in part load.Figure 3 shows a typical real world example: Pipeline operating points for different flow requireme nts are plotted into the performa nee map of the speed con trolled cen trifugal compressor used in the compressor stati on.Reciprocati ng compressors wi

54、ll automatically comply with the system pressure ratio dema nds,as long as no mecha ni cal limits (rod load power)are exceeded. Chan ges in system suction or discharge pressure will simply cause the valves to open earlier or later. The head is lowered automatically because the valves see lower pipel

55、i ne pressures on the discharge side an d/or higher pipeli ne pressures on the suct ion side. Therefore, without additi onal measures, the flow would stay roughly the same except for the impact of cha nged volumetric efficie ncy which would in crea.se, thus increasing the flow with reduced presstire

56、 ratio.The con trol challe nge lies in the adjustme nt of the flow to the system dema nds. Without additi onal adjustme nts, the flow throughput of the compressor cha nges very little with cha nged pressure ratio. Historically, pipeli nes in stalled many small compressors and adjusted flow rate by c

57、ha nging the nu mber of mach ines activated. This capacity and load could be finetuned by speed or by a number of small adjustme nts (load steps) made in the cyli nder cleara nee of a sin gle un it. As compressors have grow n, the burde n for capacity con trol has shifted to the in dividual compress

58、ors.Load control is a critical component to compressor operation. From a pipeline operation perspective, variation in station flow is required to meet pipeline delivery commitme nts, as well as impleme nt compa ny strategies for optimal operatio n (i.e., line pack in g, load an ticipati on ).From a unit perspective, load con trol invo Ives reducing unit flow (through unioaders or speed)to operate as close as possible to the design torque limit without overloading the compressor or driver The critical limits on any load map curve are rod load limits and HP/torque limits for any giv