1、離心式和往復(fù)式壓縮機(jī)的工作效率特性離心式和往復(fù)式壓縮機(jī)的工作效率特性 Rainer Kurz , Bernhard Winkelmann , 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 際壓縮過程的工作效率。難點在于正確定義適當(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

2、些邊界是恰好定義的，否則離心式和往復(fù)式壓縮機(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，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

3、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 者單位的運行在 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 (h)，總壓力(p)，溫度(T)和熵(s)，于是等熵效率s變?yōu)椋?h(p disch ,s suct )h(p suct ,T suct ) s h(p disch ,T disch )h(p suct ,T suct ) (Eq.1) 并且加上測量的穩(wěn)態(tài)質(zhì)量流 m，吸收軸功率為： p m . m h(p disch ,T di

4、sch )h(p suct ,T suct ) (Eq.2) 考慮機(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）如下： (Eq.3) 流入和流出離心式壓縮機(jī)的流量可以視為 “穩(wěn)態(tài)” 。 環(huán)境的熱交換通?？梢院雎?。 系統(tǒng)邊界的效率計算通常是用吸入和排出的噴嘴。需要確定的是，系統(tǒng)邊界要包 含所有內(nèi)部泄露途徑，尤其是從平衡活塞式或分裂墻滲漏的循環(huán)路徑。機(jī)械效率 P theor mh(p disch ,s suct ) h(p suct ,T suct ) . m，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 98%和 99%。 word 文檔 可自由復(fù)制編輯 對于往復(fù)式壓縮機(jī)，

5、理論的氣體馬力也是由 Eq.3 給出的，鑒于吸力緩沖器 上游和排力緩沖器下游的吸氣和排氣壓力脈動。往復(fù)壓縮機(jī)就其性質(zhì)而言，從臨 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） ，以及 可以自然存在的來自管線的管流量和面積管道。 對于任何一個低速或高速單位的 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。 潛在的， 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。就像離心壓 縮機(jī)一樣，傳熱就經(jīng)常被忽視。 對于積分的機(jī)器，機(jī)械效率一般取為 95%。對于可分機(jī)機(jī)械效率一般使用 97%

6、。 這些數(shù)字似乎有些樂觀， 一系列數(shù)字顯示， 往復(fù)式發(fā)動機(jī)機(jī)械損失在 8-15% 之間，往復(fù)壓縮機(jī)的在 6-12%（參考 1 往復(fù)壓縮機(jī)招致號碼：庫爾茲，R.，K.， 光布倫，2007） 。 工作環(huán)境 在這樣的情況下，當(dāng)壓縮機(jī)在一個系統(tǒng)中運行時，管道長度 Lu 上游和 Ld 下游，以及管道 pu 上游的初始壓力和管道 pe 下游的終止壓力均被視為常量，在 管道系統(tǒng)中我們有一個壓縮機(jī)運行的簡單模型（圖 1） 。 圖 1：管道段的概念模型（文獻(xiàn) 2：庫爾茲.R，M.由羅穆斯基，2006 年） 。 對于給定的，標(biāo)準(zhǔn)管線定量流動能力將在吸入階段強加壓力p s ，在壓縮機(jī) 放電區(qū)強加壓力p d 。對于給

7、定的管線，壓縮機(jī)站頭部（H s ）流（Q）關(guān)系可以 1 近似表述為H s C pTs C 3 C 4 Q2 2 1 p d k1 k 1 （Eq.4） 其中C 3 和C 4 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 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， 這意味著對于帶管道系統(tǒng)的壓縮機(jī)站，頭部所需流量揚程是 由管道系統(tǒng)規(guī)定的（圖 2） 。特別地，這一特點對于壓縮機(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

8、） ，最初的操作條件遵循恒功率分布，如頭部 流量關(guān)系如下： P m H s s const（Eq.5） H s s const1 Q 并將漸進(jìn)地達(dá)到穩(wěn)定的關(guān)系（文獻(xiàn) 3：奧海寧 S.，R.庫爾茲，2002 年） word 文檔 可自由復(fù)制編輯 在上述要求的基礎(chǔ)上， 必須控制壓縮機(jī)輸出與系統(tǒng)要求匹配。該系統(tǒng)需求的 特點是系統(tǒng)流程和系統(tǒng)頭部或壓力比的強烈關(guān)系。 管線壓縮機(jī)提供了在操作條件 經(jīng)驗下的大量變化，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，具體 的說就是如何影響效率。 離心壓縮機(jī)具有相當(dāng)大的平頭部和流程特點。 這意味著壓力比的改變對機(jī)器 的實際流程有重大的影響（文獻(xiàn) 4：庫爾茲 R.

9、，20004 年） 。對于一個恒速運行 的壓縮機(jī)， 頭部或壓力比隨著流量的增加而減少?？刂茐嚎s機(jī)內(nèi)的流程可以實現(xiàn) 壓縮機(jī)不同的運行速度。這是控制離心壓縮機(jī)最便捷的方法。兩軸燃?xì)廨啓C(jī)和變 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50%到 100%） 。 應(yīng)當(dāng)指出，被控制的值通常不是速度，但速度是間接平衡由渦輪產(chǎn)生的動力（受 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。 事實上，在過去 15 年安裝的任何離心壓縮機(jī)在管線服務(wù)方面是由調(diào)速器來 驅(qū)使的， 通常是兩軸燃?xì)廨啓C(jī)。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 單軸燃?xì)廨啓C(jī)（允許速度 90%到 100%的變化）和恒速電動

10、機(jī)。在這些裝置中， 吸節(jié)流或可變進(jìn)氣導(dǎo)葉用來提供控制方法。 圖 3：典型的管線運行點繪制成的典型離心壓縮機(jī)性能圖。 離心壓縮機(jī)的運行封套受最大允許速度限制， 最小流量 （涌） 和最大流量 （窒 息或石墻） （圖 3） 。另一個限制因素可能是可用的驅(qū)動電源。 只有最小流量需要特別注意， 因為它被定義為壓縮機(jī)的一種氣動穩(wěn)定性的極 限?？缭竭@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)。調(diào) 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。出于這個原因，幾乎所 有的現(xiàn)代壓縮機(jī)裝置都使用帶有控制閥的循環(huán)線， 當(dāng)壓縮機(jī)內(nèi)的流量趨于穩(wěn)定極 限時這種控制閥允許流量的增加。 控制系統(tǒng)不斷地監(jiān)測壓縮

11、機(jī)關(guān)系喘振線的運行 點，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，帶有開放或部分 開放循環(huán)閥的運行模式只被用于開啟和關(guān)閉階段， 或者是在混亂運行條件時的短 暫時期。 假設(shè)由公式 4 得到管線特點， 壓縮機(jī)的葉輪將在達(dá)到或接近其最大效率時被 選出來運行， 這個最大效率是由管線強加在整個系列的頭部和流量條件下的。這 可能是有一個速度（N）控制的壓縮機(jī)，因為一個壓縮機(jī)的最有效點是由一種關(guān) 系而連接的，這種關(guān)系需要大約（風(fēng)扇法方程） ： H 5 Q CH Q2 C 5(Eq.6) C 565NN2 C 6 2 為滿足上述關(guān)系的操作點，吸入氣壓Pg是（基于效率幾乎保持不變這個的 事實） ： P

12、 g C 7 H 5 Q C 533C Q C C C N(Eq.7) 7567 2C 6 正因為如此，這種力-速度關(guān)系允許動力渦輪運行達(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 最有效點運行。 然而，燃?xì)廨啓C(jī)的燃?xì)馍a(chǎn)商將在部分負(fù)荷運行時丟失一些熱效 率。 圖 3 顯示了一個典型的實際例子： 不同流動要求的管線運行點繪制成用于壓 word 文檔 可自由復(fù)制編輯 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或晚的開啟。頭部 可

13、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，流量將大致恒定除了容積效率將增加的變化， 所以降低壓力比而增加流量。 控制的挑戰(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)移到獨立壓縮機(jī)上。 負(fù)荷控制是壓縮機(jī)運行的一個關(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線投出承諾，以及

14、實施公司最佳操作（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

15、分可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。 運行運行 典型的穩(wěn)態(tài)管道運行將產(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

16、在文獻(xiàn) 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)缺點討論將變得 不準(zhǔn)確和有誤導(dǎo)性。 word 文檔 可自由復(fù)制編輯 參考

17、文獻(xiàn)： 1.庫爾茲.R.K.光布倫，2007。 “往復(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 GT2007-2708 2.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 2006-90069 3.奧海寧.S.R.庫爾茲，2002。 “兩機(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，

18、UT 原文 EfficiencyAndOperatingCharacteristicsOfCentrifugalAndReciprocating Compressors By Rainer Kurz, Bernhard Winkelmann, and Saeid iVIokhatab Reciprocating compressors and centrifugal compressors have different operating characteristics and use different eificiency definitions. This article provide

19、s guidelines for an equitable comparison, 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 diffe

20、rent load levels is evaluated. At first glance, calculating the efficiency for any type of compression seems to be straightforward: comparing the work required of an idealcompression process with the work required of an actual compression process. The difficulty is correctly defining appropriate sys

21、tem boundaries that include losses associated with the compression process. Unless these boundaries are appropriately defined, comparisons between centrifugal and reciprocating compressors become flawed. We also need to acknowledge that the efficiency definitions, even when evaluated equitably, stil

22、l dont completely answer one of the operators main concerns: What is the driver power required for the compression process?To accomplish this, mechanical losses in the compression systems need to be discussed. Trends in efficiency should also be considered over time, such as off-design conditions as

23、 they are imposed by typical pipeline operations, or the impact of operating hours and associated degradation on the compressors. The compression equipment used for pipelines involves either reciprocating compressors or centrifugal compressors. Centrifugal compressors are driven by gas turbines, or

24、by electricmotors. The gas turbines used are, in general,two-shaft engines and the electric motor drives use either variable speed motors, or variable speed gearboxes. Reciprocating compressors are either low speed integral units, which combine the gas engine and the compressor in one crank casing,o

25、r separable word 文檔 可自由復(fù)制編輯 high-speed units. The latter units operate in the 750-1,200 rpm range (1,800 rpm for smaller units) and are generally driven by electric motors, or four-stroke gas engines. Efficiency To determine the isentropic efficiency of any compression processbased on total enthalpi

26、es (h), total pressures (p), temperatures (T)and entropies (s) at suction and discharge of the compressor are measured, and the isentropic efficiency r then becomes: s h(p disch ,s suct )h(p suct ,T suct ) h(p disch ,T disch )h(p suct ,T suct ) (Eq.1) and, with measuring the steady state mass flow m

27、, the absorbed shaft power is: p m . m h(p disch ,T disch )h(p suct ,T suct ) (Eq.2) considering the mechanical efficiency r. The theoretical (isentropic) power consumption (which is the lowest possible power consumption for an adiabatic system) follows from: (Eq.3) The flow into and out of a centri

28、fugal compressor can be considered as steady state.Heat exchange with the environment is usually negligible. System boundaries for the efficiency calculations are usually the suction and discharge nozzles. It needs to be assured that the system boundaries envelope all internal leakage paths,in parti

29、cular recirculation paths fiom balance piston or division wall leakages. The mechanical efficiency r)., describing the friction losses in bearings and seals, as well as windage losses, is typically between 98 and 99%. For reciprocating compressors, theoretical gas horsepower is also given by Eq. 3,g

30、iven the suction and discharge pressure are upstream of the suction pulsation dampeners and downstream of the discharge pulsation dampeners. Reciprocating compressors, by their very nature, require manifold systems to control pulsations and provide isolation from neighboring units (both reciprocatin

31、g and centrifugal), as well as from pipeline flow meters and yard piping and can be extensive in nature.The design of manifold systems for either slow speed or high speed units uses a combination of volumes, piping lengths and pressure drop elements to create pulsation (acoustic) filters.These manif

32、old systems (filters) cause a pressure drop, and thus must be considered in efficiency calculations. Potentially, additional pressure deductions from the suction pressure would have to made to include the effects of residual pulsations. Like centrifugal compressors, heat transfer is usually neglecte

33、d. For integral machines, mechanical efficiency is generally taken as 95%. For separable machines a 97% mechanical efficiency is often used. These numbers seem to be somewhat optimistic, given the fact that a number of sources state that reciprocating engines incur between 8-15% mechanical losses an

34、d reciprocating compressors between 6-12%(Ref 1: Kurz , R., K. Brun, 2007). word 文檔 可自由復(fù)制編輯 P theor mh(p disch ,s suct ) h(p suct ,T suct ) . Operating Conditions For a situation where a compressor operates in a system with pipe of the length Lu upstream and a pipe of the length Ld downstream, and f

35、urther where the pressure at the beginning of the upstream pipe pu and the end of the downstream pipe pe are known and constant, we have a simple model of a compressor station operating in a pipeline system (Figure 1). Figure 1: Conceptual model of a pipeline segment (Ref. 2: Kurz, R., M. Lubomirsky

36、.2006). For a given, constant flow capacity Qstd the pipeline will then impose a pressure ps at the suction and pd at the discharge side of the compressor. For a given pipeline, the head (Hs)-flow (Q) relationship at the compressor station can be approximated by k1 k 1 H s C pTs 1 （Eq.4） 2 C 3 C 4 Q

37、 1 2 p d where C3 and C4 are constants (for a given pipeline geometry) describing the pressure at either ends of the pipeline, and the friction losses, respectively(Ref 2: Kurz, R., M. Lubomirsky, 2006). Among other issues, this means that for a compressor station within a pipeline system, the head

38、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 reduction in head with reduced flow, and vice versa, in a prescribed fashion. The pipeline will therefore not require a change in flow at con

39、stant head (or pressure ratio). Figure 2: Stafion Head-Flow relationship based on Eq. 4. In transient situations (for example during line packing), the operating conditions follow initially a constant power distribution, i.e. the head flow relationship follows: P m H s s const（Eq.5） H s s const1 Q a

40、nd will asymptotically approach the steady state relationship (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 relationship between system flow and system head or

41、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 compressors tend to have rather flat head vs

42、. flow characteristic. This word 文檔 可自由復(fù)制編輯 means that changes in pressure ratio have a significant effect on the actual flow 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 increasing flow. Controllin

43、g the flow through the compressor can be accomplished by varying the operating speed of the compressor This is the preferred method of controlling centrifugal compressors. Two shaft gas turbines and variable speed electric motors allow for speed variations over a wide range (usually from 40-50% to 1

44、00% of 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 power turbine (which is controlled by the fuel flow into the gas turbine) and the absorbed power of the compressor. Virtual

45、ly any centrifugal compressor installed in the past 15 years in pipeline service is driven by a variable speed driver, usually a two-shaft gas turbine. Older installations and installations in other than pipeline service sometimes use single-shaft gas turbines (which allow a speed variation from abo

46、ut 90-100% speed) and constant speed electric motors. In these installations, suction throttling or variable inlet guide vanes are used to Drovide means of control. Figure 3: Typical pipeline operating points plotted into a typical centrifugal compressor performance map. The operating envelope of a

47、centrifugal compressor is limited by the maximum allowable speed, the minimum flow (surge flow),and the maximum flow (choke or stonewall)(Figure 3). Another limiting factor may be the available driver power. Only the minimum flow requires special attention, because it is defined by an aerodynamic st

48、ability limit of the compressor Crossing this limit to lower flows will cause a flow reversal in the compressor, which can damage the compressor. Modem control systems prevent this situation by automatically opening a recycle valve. For this reason, virtually all modern compressor installations use

49、a recycle line with 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 recycle valve if necessar

50、y. For most applications, the operating mode with an open, or partially open recycle valve is only used for start-up and shutdown, or for brief periods during upset operating conditions. Assuming the pipeline characteristic derived in Eq. 4, the compressor impellers will be selected to operate at or

51、 near its best efficiency for the entire range of head and flow conditions imposed by the pipeline. This is possible with a speed (N) controlled compressor, because the best efficiency points of a compressor are connected by a relationship that requires approximately (fan law equation): H 5 Q CH Q2

52、C 5(Eq.6) C 565NN2 C 6 2 For operating points that meet the above relationship, the absorbed gas power Pg is (due to the fact that the efficiency stays approximately constant): word 文檔 可自由復(fù)制編輯 P g C 7 H 5 Q C 533C Q C C C N(Eq.7) 7567 2C 6 As it is, this power-speed relationship allows the power tur

53、bine to operate at, or very close to its optimum speed for the entire range.The typical operating scenarios in pipelines therefore allow the compressor and the power turbine to operate at its best efliciency for most of the time. The gas producer of the gas turbine will, however, lose some thermal e

54、fficiency when operated in part load. Figure 3 shows a typical real world example: Pipeline operating points for different flow requirements are plotted into the performance map of the speed controlled centrifugal compressor used in the compressor station. Reciprocating compressors will automaticall

55、y comply with the system pressure ratio demands,as long as no mechanical limits (rod load power)are exceeded. Changes 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 pipeline pressures on th

56、e discharge side and/or higher pipeline pressures on the suction side. Therefore, without additional measures, the flow would stay roughly the same except for the impact of changed volumetric efficiency which would increa.se, thus increasing the flow with reduced presstire ratio. The control challen

57、ge lies in the adjustment of the flow to the system demands. Without additional adjustments, the flow throughput of the compressor changes very little with changed pressure ratio. Historically, pipelines installed many small compressors and adjusted flow rate by changing the number of machines activ

58、ated. This capacity and load could be fine-tuned by speed or by a number of small adjustments (load steps) made in the cylinder clearance of a single unit. As compressors have grown, the burden for capacity control has shifted to the individual compressors. Load control is a critical component to co

59、mpressor operation. From a pipeline operation perspective, variation in station flow is required to meet pipeline delivery commitments, as well as implement company strategies for optimal operation (i.e., line packing, load anticipation).From a unit perspective, load control involves reducing unit flow (through unloaders or speed)to operate as close as po