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編號(hào)
無錫太湖學(xué)院
畢業(yè)設(shè)計(jì)（論文）
相關(guān)資料
題目：微型風(fēng)冷活塞式壓縮機(jī)(W-80)的設(shè)計(jì)
信機(jī) 系 機(jī)械工程及自動(dòng)化專業(yè)
學(xué) 號(hào)： 0923208
學(xué)生姓名： 顧 佳 慶
指導(dǎo)教師： 俞萍（職稱：高級(jí)工程師）
（職稱： ）
2013年5月25日
目 錄
一、畢業(yè)設(shè)計(jì)（論文）開題報(bào)告
二、畢業(yè)設(shè)計(jì)（論文）外文資料翻譯及原文
三、學(xué)生“畢業(yè)論文（論文）計(jì)劃、進(jìn)度、檢查及落實(shí)表”
四、實(shí)習(xí)鑒定表
無錫太湖學(xué)院
畢業(yè)設(shè)計(jì)（論文）
開題報(bào)告
題目：微型風(fēng)冷活塞式壓縮機(jī)(W-80)的設(shè)計(jì)
信機(jī) 系 機(jī)械工程及自動(dòng)化 專業(yè)
學(xué) 號(hào)： 0923208
學(xué)生姓名： 顧佳慶
指導(dǎo)教師： 俞萍（職稱：高級(jí)工程師）
（職稱： ）
2012年11月20日
課題來源
本課題來源于企業(yè)；
結(jié)合所學(xué)知識(shí)，老師擬定題目；
綜合大學(xué)里所學(xué)知識(shí)，將理論與實(shí)踐相互結(jié)合。
科學(xué)依據(jù)（包括課題的科學(xué)意義；國(guó)內(nèi)外研究概況、水平和發(fā)展趨勢(shì)；應(yīng)用前景等）
1、 化工、冶金、化肥、食品、醫(yī)療等眾多企業(yè)的生產(chǎn)過程需要用到氣體
壓縮機(jī)，而活塞式空氣壓縮機(jī)由于有較高的壓縮比，在高壓氣體生產(chǎn)與輸送中尚不能被其它設(shè)備所替代，是許多工程項(xiàng)目中的關(guān)鍵設(shè)備。
2、 活塞式壓縮機(jī)在圓筒形氣缸中具有一個(gè)可往復(fù)運(yùn)動(dòng)的活塞，氣缸上有控制進(jìn)、排氣的閥門，當(dāng)活塞作往復(fù)運(yùn)動(dòng)時(shí)，氣缸的容積便周期性的變化，借以實(shí)現(xiàn)氣體的吸進(jìn)、壓縮、和排出。
3、 隨著經(jīng)濟(jì)的高速發(fā)展和科學(xué)技術(shù)的不斷進(jìn)步，各種壓縮機(jī)在國(guó)民經(jīng)濟(jì)各大領(lǐng)域大顯身手，壓縮機(jī)是原基礎(chǔ)材料之一的冶金工業(yè)中極為重要的設(shè)備，又是石油化工流程中的心臟設(shè)備。車輛的制動(dòng)、船用內(nèi)燃機(jī)啟動(dòng)，航空發(fā)動(dòng)機(jī)的運(yùn)行都需要各種壓縮機(jī)，可以說壓縮機(jī)在陸?？战煌ㄟ\(yùn)輸工具中都必不可少，與人民的日常生活更是休戚相關(guān)。
4、 目前壓縮機(jī)制造業(yè)已經(jīng)發(fā)展成為機(jī)械制造工業(yè)的一個(gè)重要組成部分。
研究?jī)?nèi)容
1、 微型風(fēng)冷活塞式壓縮機(jī)的工作原理以及工作形成；
2、 微型風(fēng)冷活塞式壓縮機(jī)參數(shù)與結(jié)構(gòu)的設(shè)計(jì)；
3、 微型風(fēng)冷活塞式壓縮機(jī)設(shè)計(jì)圖紙的繪制。
擬采取的研究方法、技術(shù)路線、實(shí)驗(yàn)方案及可行性分析
研究方法：通過閱讀有關(guān)資料，文獻(xiàn)，收集篩選，整理課題研究所需的
有關(guān)數(shù)據(jù)，理論依據(jù)，綜合運(yùn)用所學(xué)理論知識(shí)研究論文課題。
技術(shù)路線：分析微型風(fēng)冷活塞式壓縮機(jī)的各個(gè)參數(shù)的取值情況，包括結(jié)
構(gòu)參數(shù)、工藝參數(shù)、熱力學(xué)參數(shù)和動(dòng)力學(xué)參數(shù)。確定各參數(shù)
的具體數(shù)值或取值區(qū)間。
可行性分析：通過對(duì)論文課題的學(xué)習(xí)研究，達(dá)到鞏固，擴(kuò)大，深化已學(xué)
理論知識(shí)，提高思考分析解決實(shí)際問題等綜合素質(zhì)的目的。
研究計(jì)劃及預(yù)期成果
1、 首先對(duì)微型風(fēng)冷活塞式壓縮機(jī)整體結(jié)構(gòu)進(jìn)行分析，對(duì)傳動(dòng)結(jié)構(gòu)進(jìn)行篩選，初步選擇達(dá)到設(shè)計(jì)要求的結(jié)構(gòu)方案；
2、 對(duì)壓縮機(jī)的熱力部分及動(dòng)力部分進(jìn)行計(jì)算，通過壓縮機(jī)機(jī)構(gòu)的分析計(jì)算可提高其自身的精度；
3、 對(duì)微型風(fēng)冷活塞式壓縮機(jī)的主要零件進(jìn)行強(qiáng)度校核，提高機(jī)構(gòu)穩(wěn)定性，穩(wěn)定性。
特色或創(chuàng)新之處
通過對(duì)微型風(fēng)冷活塞式壓縮機(jī)的設(shè)計(jì)及計(jì)算，形成一整套現(xiàn)代的設(shè)計(jì)方法，對(duì)理論和實(shí)踐的結(jié)合，起到整體的規(guī)劃的作用，達(dá)到降低損耗提高效率，優(yōu)化結(jié)構(gòu)設(shè)計(jì)方便使用。
已具備的條件和尚需解決的問題
已具備的條件：擁有機(jī)械設(shè)計(jì)手冊(cè)等參考資料及文獻(xiàn)；對(duì)活塞式壓縮機(jī)進(jìn)行直觀的了解與認(rèn)識(shí)，對(duì)所學(xué)的機(jī)械基礎(chǔ)知識(shí)有較好的掌握；能熟練運(yùn)用CAXA制圖軟件，提高作圖效率。
尚需解決的問題：對(duì)于微型風(fēng)冷活塞式壓縮機(jī)的工作原理不是非常清楚
和熟悉，缺乏設(shè)計(jì)經(jīng)驗(yàn)。
指導(dǎo)教師意見
指導(dǎo)教師簽名：
年 月 日
教研室（學(xué)科組、研究所）意見
教研室主任簽名：
年 月 日
系意見
主管領(lǐng)導(dǎo)簽名：
年 月 日
無錫太湖學(xué)院
畢業(yè)設(shè)計(jì)（論文）外文資料翻譯
信機(jī) 系 機(jī)械工程及自動(dòng)化 專業(yè)
院 （系）： 信 機(jī) 系
專 業(yè)： 機(jī)械工程及自動(dòng)化
班 級(jí)： 機(jī)械95班
姓 名： 顧 佳 慶
學(xué) 號(hào)： 0923208
外文出處： 機(jī)械專業(yè)英語(yǔ)教程
附 件： 1.譯文；2.原文；3.評(píng)分表
2013年5月25日
英文原文
Efficiency 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 don't completely answer one of the operator's 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.
Efficiency
To 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:
(Eq.1)
and, with measuring the steady state mass flow m, the absorbed shaft power is:
(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 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 from 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. Bun, 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 by
（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:
（Eq.5）
and 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):
(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):
(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 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 efficiency 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 automatically 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 the 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 increase, thus increasing the flow with reduced presstire ratio.
The control challenge 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 activated. 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 compressor 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 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 given station suction and discharge pressure. Gas control generally will establish the units within a station that must be operated to achieve pipeline flow targets. Local unit control will establish load step or speed requirements to limit rod loads or achieve torque control.
The common methods of changing flow rate are to change speed, change clearance, or de-activate a cylinder-end (hold the suction valve open). Another method is an infinite-step unloader, which delays suction valve closure to reduce volumetric efficiency. Further, part of the flow can be recycled or the suction pressure can be throttled thus reducing the mass flow while keeping the volumetric flow into the compressor approximately constant.
Control strategies for compressors should allow automation, and be adjusted easily during the operation of the compressor .In particular, strategies that require design modifications to the compressor (for example: re-wheeling of a centrifugal compressor, changing cylinder bore, or adding fixed clearances for a reciprocating compressor)are not considered here. It should be noted that with reciprocating compressors, a key control requirement is to not overload the driver or to exceed mechanical limits.
Operation
The typical steady state pipeline operation will yield an efficiency behavior as outlined in Figure 4. This figure is the result of evaluating the compressor efficiency along a pipeline steady state operating characteristic. Both compressors would be sized to achieve their best efficiency at 100% flow, while allowing for 10% flow above the design flow. Different mechanical efficiencies have not been considered for this comparison.
The reciprocating compressor efficiency is derived n-on valve efficiency measurements in Ref 5 (Null, M., W. Couch, 2003) with compression efficiency and losses due to pulsation attenuation devices added. The efficiencies are achievable with low speed compressors. High speed reciprocating compressors may be lower in efficiency.
Figure 4: Compressor Efficiency at different flow rates based on operation along a steady state pipeline characteristic.
Figure 4 shows the impact of the increased valve losses at lower pressure ratio and lower flow for reciprocating machines, while the efficiency of the centrifugal compressor stays more or less constant.
Conclusions
Efficiency definitions and comparison between different types of compressors require close attention to the definition of the boundary conditions for which the efficiencies are defined as well as the operating scenario in which they are employed. The mechanical efficiency plays an important role when efficiency values are used to calculate power consumption. If these definitions are not considered, discussions of relative merits of different systems become inaccurate and misleading.
REFERENCES
1 Kurz . R.. K. Burn. 2007. " Efficiency Definition and Load Management for Reciprocating and Centrifugal Compressors," ASME Paper GT2OO7-27O81.
2 Kurz. R., M. Lubomirsky, 2006. "Asymttietric Solution for Compressor Station Spare Capacity."ASMt: Paper 2006-90069.
3 Ohanian. S.. R. Kurz. 2002, "Series or Parallel Arrangement in a Two-Unit Compressor Station." Trans.ASME Jeng for GT and Power. Vol.124.
4 Kurz. R.. 2004. "The Physies of Centrifugal Compressor Performance." Pipeline Simulation Interest Group. Palm Springs. CA.
5 Noral, M.. W. Couch. 2003, "Performance and Endurance Tests of Six Mainline Compressor Valves in Natural Gas Compression Service." Gas Machinery Conference. Salt Lake City. UT.
中文原文
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(gè)公平的比較準(zhǔn)則，得到了對(duì)于兩種類型機(jī)器普遍適用的效率定義。這個(gè)比較基于用戶最感興趣的要求提出的。此外，對(duì)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(píng)估。
乍一看，計(jì)算任何類型的壓縮效率看似是很簡(jiǎn)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(píng)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(dòng)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(shí)間的推移效率趨勢(shì)也應(yīng)被考慮，如非設(shè)計(jì)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(shí)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)廨啓C(jī)或者是電動(dòng)馬達(dá)來驅(qū)動(dòng)。所用的燃?xì)廨啓C(jī)，總的來說，是兩軸發(fā)動(dòng)機(jī)，電動(dòng)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(dòng)機(jī)和壓縮機(jī)在一個(gè)曲柄套管內(nèi)。后者單位的運(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(dòng)馬達(dá)或者四沖程燃?xì)獍l(fā)動(dòng)機(jī)來驅(qū)動(dòng)。
效率
要確定任何壓縮過程的等熵效率，就要基于測(cè)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)，于是等熵效率變?yōu)椋? (Eq.1)
并且加上測(cè)量的穩(wěn)態(tài)質(zhì)量流m，吸收軸功率為：
(Eq.2)
考慮機(jī)械效率。
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(jì)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，在描述軸承和密封件的摩擦損失以及風(fēng)阻損失時(shí)可以達(dá)到98%和99%。
對(duì)于往復(fù)式壓縮機(jī)，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(dòng)。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(dòng)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(duì)于任何一個(gè)低速或高速單位的歧管