离心式和往复式压缩机的工作效率特性【中文4160字】【中英文WORD】.zip

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【中文【中文 4160 字】字】离心式和往复式压缩机的工作效率特性离心式和往复式压缩机的工作效率特性 Rainer Kurz,Bernhard Winkelmann,and Saeid Mokhatab往复式压缩机和离心式压缩机具有不同的工作特性,而且关于效率的定义也不同。本文提供了一个公平的比较准则,得到了对于两种类型机器普遍适用的效率定义。这个比较基于用户最感兴趣的要求提出的。此外,对于管道的工作环境影响和在不同负载水平的影响给出了评估。乍一看,计算任何类型的压缩效率看似是很简单的:比较理想压缩过程和实际压缩过程的工作效率。难点在于正确定义适当的系统边界,包括与之相关的压缩过程的损失。除非这些边界是恰好定义的,否则离心式和往复式压缩机的比较就变得有缺陷了。我们也需要承认,效率的定义,甚至是在评估公平的情况下,仍不能完全回应操作员的主要关心问题:压缩过程所需的驱动力量是什么?要做到这一点,就需要讨论在压缩过程中的机械损失。随着时间的推移效率趋势也应被考虑,如非设计条件,它们是由专业的流水线规定,或者是受压缩机的工作时间和自身退化的影响。管道使用的压缩设备涉及到往复式和离心式压缩机。离心式压缩机用燃气轮机或者是电动马达来驱动。所用的燃气轮机,总的来说,是两轴发动机,电动马达使用的是变速马达或者变速齿轮箱。往复压缩机是低速整体单位或者是可分的“高速”单位,其中低速整体单位是燃气发动机和压缩机在一个曲柄套管内。后者单位的运行在 750-1,200rpm 范围内(1,800rpm 是更小的单位)并且通常都是由电动马达或者四冲程燃气发动机来驱动。效率效率要确定任何压缩过程的等熵效率,就要基于测量的压缩机吸入和排出的总焓(h),总压力(p),温度(T)和熵(s),于是等熵效率s变为:),(),(),(),(suctsuctdischdischsuctsuctsuctdischsTphTphTphsph (Eq.1)并且加上测量的稳态质量流 m,吸收轴功率为:),(),(.suctsuctdischdischmTphTphmp (Eq.2)考虑机械效率m。理论(熵)功耗(这是绝热系统可能出现的最低功耗)如下:),(),(.suctsuctsuctdischtheorTphsphmP (Eq.3)流入和流出离心式压缩机的流量可以视为“稳态”。环境的热交换通常可以忽略。系统边界的效率计算通常是用吸入和排出的喷嘴。需要确定的是,系统边界要包含所有内部泄露途径,尤其是从平衡活塞式或分裂墙渗漏的循环路径。机械效率m,在描述轴承和密封件的摩擦损失以及风阻损失时可以达到 98%和 99%。对于往复式压缩机,理论的气体马力也是由 Eq.3 给出的,鉴于吸力缓冲器上游和排力缓冲器下游的吸气和排气压力脉动。往复压缩机就其性质而言,从临近单位需要多方面的系统来控制脉动和提供隔离(包括往复式和离心式),以及可以自然存在的来自管线的管流量和面积管道。对于任何一个低速或高速单位的歧管系统设计,使用了卷相结合,管道长度和压力降元素来创造脉动(声波)滤波器。这些歧管系统(过滤器)引起压力下降,因此必须在效率计算时考虑到。潜在的,从吸气压力扣除的额外压力不得不包含进残余脉动的影响。就像离心压缩机一样,传热就经常被忽视。对于积分的机器,机械效率一般取为 95%。对于可分机机械效率一般使用97%。这些数字似乎有些乐观,一系列数字显示,往复式发动机机械损失在 8-15%之间,往复压缩机的在 6-12%(参考 1 往复压缩机招致号码:库尔兹,R.,K.,光布伦,2007)。工作环境在这样的情况下,当压缩机在一个系统中运行时,管道长度 Lu 上游和 Ld下游,以及管道 pu 上游的初始压力和管道 pe 下游的终止压力均被视为常量,在管道系统中我们有一个压缩机运行的简单模型(图 1)。图 1:管道段的概念模型(文献 2:库尔兹.R,M.由罗穆斯基,2006 年)。对于给定的,标准管线定量流动能力将在吸入阶段强加压力sp,在压缩机放电区强加压力dp。对于给定的管线,压缩机站头部(sH)流(Q)关系可以近似表述为 11112243skkdsppQCCTCH (Eq.4)其中3C和4C是常数(对于一个给定的管道几何)分别描述了管道两边的压力和摩擦损失(文献 2:库尔兹.R,M.由罗穆斯基,2006 年)。除去其他问题,这意味着对于带管道系统的压缩机站,头部所需流量扬程是由管道系统规定的(图 2)。特别地,这一特点对于压缩机需要的能力允许头部减量,按照规定的方式反之亦然。管道因此将不需要改变头部的流量恒定(或压力比)。图 2:建立在 4 式上的机头流量关系。在短暂的情况下(如包装其间),最初的操作条件遵循恒功率分布,如头部流量关系如下:constHPssm (Eq.5)QconstHss1并将渐进地达到稳定的关系(文献 3:奥海宁 S.,R.库尔兹,2002 年)在上述要求的基础上,必须控制压缩机输出与系统要求匹配。该系统需求的特点是系统流程和系统头部或压力比的强烈关系。管线压缩机提供了在操作条件经验下的大量变化,一个重要问题就是如何使压缩机适应这样变化的条件,具体的说就是如何影响效率。离心压缩机具有相当大的平头部和流程特点。这意味着压力比的改变对机器的实际流程有重大的影响(文献 4:库尔兹 R.,20004 年)。对于一个恒速运行的压缩机,头部或压力比随着流量的增加而减少。控制压缩机内的流程可以实现压缩机不同的运行速度。这是控制离心压缩机最便捷的方法。两轴燃气轮机和变速电机允许大范围的速度变化(通常是最大速度或更多的 40%或 50%到 100%)。应当指出,被控制的值通常不是速度,但速度是间接平衡由涡轮产生的动力(受进入燃气轮机燃油流量控制)和压缩机的吸收功率。事实上,在过去 15 年安装的任何离心压缩机在管线服务方面是由调速器来驱使的,通常是两轴燃气轮机。年长的设施和服务设施在其他管线服务有时使用单轴燃气轮机(允许速度 90%到 100%的变化)和恒速电动机。在这些装置中,吸节流或可变进气导叶用来提供控制方法。图 3:典型的管线运行点绘制成的典型离心压缩机性能图。离心压缩机的运行封套受最大允许速度限制,最小流量(涌)和最大流量(窒息或石墙)(图 3)。另一个限制因素可能是可用的驱动电源。只有最小流量需要特别注意,因为它被定义为压缩机的一种气动稳定性的极限。跨越这个限制以降低流动将导致压缩机流动逆转,这可能会损坏压缩机。调制解调器控制系统通过打开一个循环阀来控制这种情况。出于这个原因,几乎所有的现代压缩机装置都使用带有控制阀的循环线,当压缩机内的流量趋于稳定极限时这种控制阀允许流量的增加。控制系统不断地监测压缩机关系喘振线的运行点,并且有必要的话自动地开关循环阀。对于大多数应用来说,带有开放或部分开放循环阀的运行模式只被用于开启和关闭阶段,或者是在混乱运行条件时的短暂时期。假设由公式 4 得到管线特点,压缩机的叶轮将在达到或接近其最大效率时被选出来运行,这个最大效率是由管线强加在整个系列的头部和流量条件下的。这可能是有一个速度(N)控制的压缩机,因为一个压缩机的最有效点是由一种关系而连接的,这种关系需要大约(风扇法方程):525CNH 6CNQ 26525CCQH (Eq.6)为满足上述关系的操作点,吸入气压gP是(基于效率几乎保持不变这个的事实):37653726557gNCCCQCCCQHCP (Eq.7)正因为如此,这种力-速度关系允许动力涡轮运行达到或非常接近其整个范围的理想速度。管线中典型的运行方案允许压缩机和动力涡轮在大多数时间里在最有效点运行。然而,燃气轮机的燃气生产商将在部分负荷运行时丢失一些热效率。图 3 显示了一个典型的实际例子:不同流动要求的管线运行点绘制成用于压缩机站中的速度控制离心压缩机性能图。往复压缩机将自动服从系统压力比的需求,只要没有超出机械的限制条件(杆负载功率)。系统吸排气压力的改变将仅能引起阀门或早或晚的开启。头部可以自动下降因为阀门可以降低排气端的管线压力和/或吸入端更高的管线压力。因此,如果没有额外的措施,流量将大致恒定除了容积效率将增加的变化,所以降低压力比而增加流量。控制的挑战存在于系统要求的流量调整。如果没有额外的调整,随着压力比的变化,压缩机流量的改变微乎其微。从历史上看,通过改变激活机器的数量使管线安装许多小的压缩机和调整流量。这个容量和负荷可通过速度调谐,或者通过一个单一单元的缸间隙中的许多小调整(加载步骤)来调谐。随着压缩机的发展,控制容量的负担转移到独立压缩机上。负荷控制是压缩机运行的一个关键组成部分。从管线操作角度来看,在机组中流量变化要符合管线投出承诺,以及实施公司最佳操作(例如,线包装,负载预期)。从一个单元的角度来看,负荷控制包含降低单元流量(通过卸载或速度)使操作尽可能的贴近设计扭矩限制,并在压缩机或驱动程序没有超载的情况下进行。对于任何给定的机组入口和出口压力,在任何负荷图曲线上的关键限制都是杆负荷限制和马力/扭矩限制。瓦斯控制通常会建立在一个机组的单元上,而这个机组运行必须达到管线流量目标。地方单元控制将建立负载步骤或速度要求来限制杆负荷或达到扭矩控制。改变流量的常用方法是改变速度,改变间隙,或取消激活缸头(保持进口阀开启)。另一种方法是卸载无限步骤,从而延缓吸气阀封闭以减少容积效率。此外,流程的一部分可以回收或吸气压力可以节流从而降低质量流量,同时保持进入压缩机的容积流量基本不间断。压缩机控制策略应该能够实现自动化,并在压缩机运行期间能够简便地调整。特别地,压缩机设计修改的战略需求(如:离心压缩机重新旋转,改变缸径,或给往复压缩机添加固定间隙)在这里不被考虑。需要指出的是,对于往复式压缩机一个关键的控制要求是不超载驱动或超过机械限制。运行运行典型的稳态管道运行将产生图 4 所示的一个有效行为。该图是评估沿管道稳定运行特征状态压缩机效率的结果。大中型压缩机都将达到 100%流量的最佳效率,并允许超出设计流量的 10%。不同的机械效率并没有考虑这种对比。往复压缩机效率在文献 5 中被推导出,从增加的阀门效率测量与压缩效率和造成的损失脉动衰减器。低速压缩机的效率是可以实现的。高速往复压缩机在效率上可能比较低。图 4:以稳态管线特性运行为基础的在不同流量率的压缩机效率。图 4 显示在较低压力比下增加的阀门损失的影响和往复机器的较低流量,而离心压缩机的效率几乎保持常量。结论结论不同型号压缩机间的效率定义和对比需要密切关注边界条件的定义,对于这样的边界条件,效率和受用的运行发展趋势同时被定义。当效率值用来计算功耗时机械效率具有重要作用。如果不考虑这些定义,不同系统的优缺点讨论将变得不准确和有误导性。参考文献:1.库尔兹.R.K.光布伦,2007。“往复和离心压缩机的效率定义和负荷管理”美国机械工程师协会 文章 GT2007-27082.库尔兹.R,M.由罗穆斯基,2006。“不对称接压缩机站闲置产能”。美国机械工程师协会 文章 2006-900693.奥海宁.S.R.库尔兹,2002。“两机压缩机站的系列或平行排列”。反式。美国机械工程师协会,第 124 栏4.库尔兹.R,2004。“离心压缩机性能的物理”。管道仿真利益集团。棕榈泉,加利福尼亚5.米.瓦特沙发,2003。“天然气压缩服务六主线压缩机阀门的性能和耐用性试验”。天然气机械会议。盐湖城,UTEfficiency And Operating Characteristics Of Centrifugal And Reciprocating 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 provides 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 different 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 ideal compression process with the work required of an actual compression process.The difficulty is correctly defining appropriate system 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,still 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 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 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,or separable 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.EfficiencyTo determine the isentropic efficiency of any compression process based on total enthalpies(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:),(),(),(),(suctsuctdischdischsuctsuctsuctdischsTphTphTphsph (Eq.1)and,with measuring the steady state mass flow m,the absorbed shaft power is:),(),(.suctsuctdischdischmTphTphmp (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:),(),(.suctsuctsuctdischtheorTphsphmP (Eq.3)The flow into and out of a centrifugal 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 particular 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,given 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 reciprocating 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 manifold 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 neglected.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 and reciprocating compressors between 6-12%(Ref 1:Kurz,R.,K.Brun,2007).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 further 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.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 by11112243skkdsppQCCTCH (Eq.4)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 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 constant 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:constHPssm (Eq.5)QconstHss1and 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 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.flow characteristic.This 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.Controlling 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 100%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.Virtually 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 about 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 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 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 control systems prevent this situation by automatically opening a recycle valve.For this reason,virtually all modern compressor installations use 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 necessary.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 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):525CNH 6CNQ 26525CCQH (Eq.6)For operating points that meet the above relationship,the absorbed gas power Pg is(due to the fact that the efficiency stays approximately constant):37653726557gNCCCQCCCQHCP (Eq.7)As it is,this power-speed relationship allows the power turbine to operate at,or very close to its optimum speed for the entire range.The typical operating scenarios in pipelines therefore allow the compresso
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