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ПЕРСПЕКТИВНЫЕ ДВИГАТЕЛИ ВНУТРЕННЕГО СГОРАНИЯ
УДК DOI: 10.30977/АТ.2219-8342.2018.43.0.5
COMPUTATIONAL AND EXPERIMENTAL DETERMINATION OF ENERGY LOSS OF THE OPERATING FLUID IN THE INTAKE SYSTEM OF THE AUTOMOBILE PISTON PNEUMATIC ENGINE USING THE EXERGY METHOD
A. Voronkov1, A. Charchenko1, I. Nikitchenko1, Ye. Novikova1, E. Teslenko1, A. Nazarov1, 1Kharkiv Nathional Automobile Highway University
Abstract. The application exergy method of the thermodynamic analysis for determining the energy loss of the operating fluid in the intake system of the automobile piston pneumatic engine using experimental data of its engine test rigs has been considered. The analysis of the performed exergy calculational and experimental research on the evaluation of energy loss in the intake system of the engine for current modes of its operation has been presented. The sample of a piston pneumatic engine for a combined automobile power unit created at the ICE Department of KhNAHU requires further improving and developing especially such components and systems as those connected to the burn process. A compressed air intake system is one of them. Compressed air is not only some working fluid for a pneumatic engine but the energy source for its intake system. The energy loss factor like reliability and durability is referred to the performance and efficiency factors. The object of the study is a pneumatic engine created by converting a gasoline four-cylinder four-stroke ICE at the ICE Department of KhNAHU. The experimental research was carried out by laboratory engine test rigs of the pneumatic engine with speed characterization provided that two thermodynamic parameters of compressed air at the intake were kept constant for each characteristic: pressure ps = idem and temperature Ts = 293 K. In each test mode all external parameters of the pneumatic engine were recorded and indicator diagrams of the first cylinder were taken. Each speed characteristic consisted of 6-8 modes ps= idem and Ts= 293 K = idem with changing the rotational rate of the crankshaft n, rpm, from minimally stable (about n=200 rpm) to maximum possible n = 1000 ± 50 rpm. It should be noted that there is an operating fluid pressure drop in the intake system chamber due to available inconvertibilities: throttling in the test port, hydraulic loss along the intake port as a result of fluid friction, swirls and other gas-dynamic phenomena. The exergy loss of the operating fluid enthalpy in the intake system Deil=Em-EHailwith the intake pressure ps = 0.5 MPais 13.3 kW, with the intake pressure ps = 0.7 MPa is 15.0 kW, with the intake pressure ps = 0.9 MPa is 16.9 kW, with the intake pressure ps = 1.1 MPa is 19.7 kW.
Key words: piston pneumatic engine, intake system, energy loss, enthalpy exergy, maximum efficient power modes.
Introduction
The sample of a piston pneumatic engine for a combined automobile power unit created at the ICE Department of KhNAHU requires further improving and developing especially such components and systems as those connected to the burn process. A compressed air intake system is one of them. Compressed air is not only some working fluid for a pneumatic engine but the energy source for its intake system. The energy loss factor like reliability and durability is referred to the performance and efficiency factors.
The method of determining the energy loss of the compressed air in the intake system of the pneumatic engine using the experimental data
[1] and well-known exergy method [2, 3] has been considered in the paper.
Purpose and Tasks
The purpose of this computational and experimental study is to develop a technique for applying the well-known exergy method to determine the energy loss of the operating fluid of the piston pneumatic engine in the intake system and the energy loss of operating fluid (compressed air) as its enthalpy exergy loss in the intake system of the piston pneumatic four-cylinder engine with intake operating pressures ps = 0.5-1.1 MPa at the modes of the maximum possible efficient power Nemax at given pressure levels ps according to speed characteristics
ps = idem with in take constant temperature Ts= 293 K.
Analysis of Publications
Nowadays there are many publications devoted to the fundamentals of the exergy analysis and its application in various fields of technology. Last decade of the twentieth century the work on creation of a mathematical apparatus of exergy analysis in solving optimization problems in the field of improving currentequipment and designing new equipment and technologywas completed. The fundamentals of exergy analysis are mostly generalized and fully described in the works of V. M. Brodyansky [2], G. N. Kostenko [3], A. R. Kotin [4], G. D. Rem [5]. The issues of the exergy method application in the field of heat engineering are studied by A. I. Andryu-shchenko [6 and others].
Almost all the studies on the exergy method published since 1964 are referred to the exergy applications in heat engineering, chemical technology, metallurgy, cryogenic engineering and economics. The discussions about the necessity of this section in thermodynamics and correctness of its main principles were left in past. In the 21st century the exergy method got known to many specialists of various fields of technology. They considered it as a convenient tool for solving their specific tasks. The main advantages of the exergy method of thermodynamic analysis are its simplicity and
This process is not adjustable. It is performed by a structurally specified period of filling the cylinder with the incoming charge of compressed air. The process of filling lasts up to 58 degrees of crankshaft turn: three degrees before-
universality [2]. It is possible to determine both absolute and relative values of energy loss due to inconvertibilities for any process or its part using the exergy balance.
So in the recently published work [7] the simplicity and accuracy of determining the energy loss of the operating fluid in case of its inconvertible leakages using the exergy method were presented. This paper shows determiningthe dependence of the exergy characteristic - enthalpy exergy loss due to external conditions: pressure and temperature of the supplied compressed air at the intakeof the power unit ps and Ts were found using experimental data and theoretical dependences between the burn process parameters.
Object of Study
The object of the study is a piston pneumatic engine with specifications presented in Table 1. The considered pneumatic engine was created at the ICE Department of KhNAHU by converting a gasoline four-cylinder four-stroke ICE. The experience of this conversion is described in detail in [1].
In accordance with the aboveformulated purposethe specific object of the study in the experiment and calculations is the intake system of the pneumatic engineprototype sample. Each cylinder of the engine has the same intake system through which the process of filling the cylinders with the incoming charge of compressed air is carried out.
top dead centre (TDC) and 55 degrees after TDC.
The cylinder in take system includesan in take port of circular cross-section with a diameter of 20 mm and a length of 800 mm, two
Table 1 - Specificationsof a piston pneumatic engine with spool-type air distribution -prototype sample designed and created at the ICE Department of KhNAHU in 2008
Parameter Notation Dimensions Numeric value
Cylinder diameter D mm 76
Piston stroke S mm 66
Number of cylinders Z - 4
Included angle of cylinder block Y degree 90
Cylinder displacement Vh dm3 0.2994
Volume ofclearance space Vo dm3 0.280
Relativeclearance volume E0=V0/Vh Eo - 0.935
Filling volume Vi dm3 0.0862
Filling rate Ei=Vi/V Ei - 0.288
Volume of reverse compression start V3 dm3 0.147
Degree of reverse compression E3=V3/Vh E3 - 0.491
Fullflowareaattheintakeline /^max J s m2i04 3.14
Fullflowareaattheexhaustline rmax jt m2i04 3.14
Adiabatic exponent of dry air к - 1.40
Adiabatic exponent of humid air, relative humidity ^=100 % [9] kp - 1.32
test ports of the same cross-section: the first is on the outer cylindrical rotor surface of the spool-type air distributor and the second is on the inner cylindrical stator surface at the beginning of the intake port. Test ports are partially or fully matched during the filling. Their combined flow area varies from zero to a maximum that is equal to the area of an open port.
The volume of the internal chamber of the intake system alongside with the object above the piston space at the top dead centre is so-called dead volume that increases the flow density of compressed air ge, kg / (kW/h).
Experimental Techniques
The experimental study was carried out by laboratory engine test rigs of the pneumatic engine with speed characterization provided that two thermodynamic parameters of compressed air at the intake were kept constant for each characteristic: pressure ps = idem and temperature Ts = 293 K. In each test mode all external parameters of the pneumatic engine were recorded and indicator diagrams of the first cylinder were taken. Each speed characteristic consisted of 6-8 modes ps= idem and Ts= 293 K = idem with changing the rotational rate of the crankshaft n, rpm, from minimally stable (about n=200 rpm) to maximum possible n = 1000 ± 50 rpm. The method of pneumatic engine testing to take the speed characteristics of the pneumatic engine is described in detail in [1].The whole experiment comprised four speed characteristics for intake pressures of the compressed air (being the subject of the greatest scientific and practical interest), ps = 0.5; 0.7; 0.9 and 1.1 MPa, at constant temperature Ts = 293 K.
Then using the obtained experimental data we constructed the dependences graphs of the efficient power Ne, kW, on the rotational speed of the shaftn, rpm and determined the maxima of the functions Ne (n) -the modes of the indicated power of the pneumatic engine for the corresponding pressure level ps at the temperature Ts, corresponding to the ambient temperature Toc = 293 K.
The operating fluid energy loss in the intake system of the pneumatic engine was determined only for these four modes Nemax mentioned above (being of exceptional interest).
Results of the Computational and
Experimental Study and Their Analysis Using the Exergy Method
The main results of this experimental and computational study are presented in Table 2
and fig. 1. They are the results of the second stage of the study, the starting points for which are the results of the first stage devoted to the determining the operating fluid energy loss at the very beginning of the working chamber of the pneumatic engine- in the spool-type air distributor due to leakage of incoming compressed air Dym, kW [7].
Thus the starting point for the second stage of the study is the level of exergy of the operating fluid enthalpy at the beginning of the intake process equal to the enthalpy exergy of Es, kW, supplied to the engine, minus the energy loss for leakage.
EBn=Es - Dym, kW. (1)
After that we defined the level of exergy of the operating fluid enthalpy at the end of the intake, i.e. at the end of the filling process, EHan, kW, that was dependant on the thermodynamic parameters of the air condition at that moment, primarily on its pressure pHan and temperature T
J Han-
It should be noted that there is an operating fluid pressure drop in the intake system chamber due to available inconvertibilities: throttling in the test port, hydraulic loss along the intake port as a result of fluid friction, swirls and other gas-dynamic phenomena.
The analysis of the obtained indicator diagrams during the engine test rigs allowed to determine the cylinder pressure values while filling. Table 2 shows the pressure values in the working chamber at various levels of compressed air intake pressure ps for maximum power modes according to speed characteristics. The thermodynamic parameters (enthalpy and entropy) were determined according to the level of air pressure and temperatureusing tables [8]. The performance was calculated as the specific enthalpy exergy eHan, kJ/kg, and the mass for the gas flow EHan, kW.
It should be noted thatchangingtemperature of the operating fluid while filling provided that the compressed air is supplied into the engine at ambient temperature Ts= Toc = 293 K may be neglectedon the basis of little heat exchange with the environment due to the absence of temperature difference and extremely short filling process (milliseconds), absence of expansion and contraction processes, and present throttling process occurred with a constant gas enthalpycan cause as known [10] a Joule-Thompson effect not exceeding the temperature drop by more than 1 K, i. e., within the accuracy of temperature measuring.
All this allows us to calculate the exergy of the operating fluid enthalpy at the end of the filling process according to its pressurepHan, MPa and temperature THan = T0.c = 293 K taking into account known dependences[2 h 5]:
- specific exergy
eHan (hHan — ho.c.) — To.c.(SHan — So.c), kJ/kg (2)
- air flow at the end of the filling process
EHan Gno^e3 eHan, (3)
where Gnone3= G - GyT, kg/s - mass second compressed airflowefficiently used; G, kg/s -compressed air flow supplied into the pneumatic engine; GyT, kg/s -experimentally found mass
Table 2 - Experimental data on the heat loss of the operating fluid in the intake system of the piston pneumatic engine during its operation according to speed characteristics ps = idem and constant intaketemperature of
compressed air Ts= 293 K at maximum power modes N™* , according to [1]
NameofValue Nota- Dimen- Pressure Ps, MPa
tion sion 0.5 0.7 0.9 1.1
Supplied (spent) amount of compressed air G Kg/s 0.0339 0.0656 0.1000 0.1583
Supplied (spent) thermal energy of the operating H kW 9.91 19.17 29.17 46.08
fluid in the form of its enthalpy Hs=G-hs % 100 100 100 100
Supplied (spent) operative heat energy - Е., kW 4.59 10.69 18.77 31.72
enthalpy energy % of Еs 100 100 100 100
The content of non-operative heat - enthalpy В.s kW 5.32 8.48 10.36 14.36
energy in the supplied heat (in the compressed air enthalpy) % of Нs 53.7 44.2 35.5 31.2
Efficient consumption of compressed air Gnone3 G GyT Gnone3 Kg/s 0.0324 0.0577 0.0860 0.1396
Specific enthalpy of compressed air under the conditions at the intake of the pneumatic engine h Kj/Kg 292.3 292.4 291.3 290.0
The specific entropy of compressed air under the conditions at the intake of the pneumatic Ss Kj/KgK 6.382 6.288 6.213 6.156
engine
Air specific enthalpy according to environmental conditions h "о.с. Kj/Kg 293.2 293.2 293.2 293.2
Air specific entropy according to environmental conditions So.c. Kj/KgK 6.847 6.847 6.847 6.847
Specific exergy of air enthalpy under the terms of pneumatic engine intake es Kj/Kg 135.4 163.0 187.7 200.3
The pressure in the working chamber at the end Р 1 вп МПа 0.395 0.495 0.580 0.640
of the filling process (intake)
The specific exergy of the operating fluid enthalpy at the end of filling ^нап Kj/Kg 116.5 135.3 150.7 155.6
The exergy of the operating fluid enthalpy at the Е kW 4.38 9.40 16.14 27.98
beginning of the intake EBn=Es - Dym, kW % of Еs 95.4 87.9 86.0 88.2
The specific exergy of the operating fluid enthalpy at the end of the filling process ^нап Kj/Kg 116.5 135.3 150.7 155.6
The exergy of the operating fluid enthalpy at the Е -^нап kW 3.77 7.80 21.95 21.72
end of filling EHan= Gn0ne3 * enan, KW % of Еs 82.1 73.0 69.0 68.2
Exergy loss of the operating fluid enthalpy in DBn kW 13.3 15.0 16.9 19.7
the intake system Dвn=Eвn-Eнan, kW % of Еs 116.5 135.3 150.7 155.6
flow of compressed air for leakage inthe given pneumatic engine, depending on its parameters at the intake; hHan, h0.c., kJ/kg -air specific enthalpy respectively for the given pressure pHan and temperature THan, and for ambient parameters poc and T0.c; T0.c, K - ambient temperature; SHan, S0.c.; kJ/(kg K) -air specific entropy respectively in terms of pHan, THan and in terms of poc and T0.c
The exergy loss of the operating fluid enthalpy in the intake system is equal to the difference in its exergy levels at the beginning and at the end of the process
DBn = Em - EHan, KW. (4)
0,5 0,7 ps ,МПа 1,1
Fig. 1. Diagrams of absolute DBn, kW, and relative DBn, % changes from Es, exergy loss of the operating fluid enthalpy due to reducing gas-dynamic loss of the pressure in the intake system during the filling while operating the pneumaticengine at maximum power Nemax with speed characteristics dependant on the level of pressure ps, MPa
Fig. 1 represents the increase of the energy loss of the operating fluid in the form of its enthalpy exergy loss in proportion to the increase of the supplied compressed air pressure at the engine intake ps not linearly, but by some increasing intensity. This is quite logical and is explained by the fact that with an increase in the pressure ps of the crankshaft rotational speed n, at which the maximum power Neis achieved, and shifted towards higher frequencies. Definitely this issue, first, l eads to increasing the hydraulic resistance in the intake and exhaust air ports.
Conclusion
A technique for applying the well-known exergy method to determine the energy loss of the operating fluid of the piston pneumatic engine in the intake system with intake operating pressures ps = 0.5-1.1 MPa at the modes of the maximum possible efficient power Nemax at given pressure levels ps according to speed characteristics ps = idem withintake constant temperature Ts= 293 K has been developed.
The scientific significance of this work is the development of a technique for applying the well-known exergy method to determining the energy loss of the operating fluid of the piston pneumatic engine in the intake system.
The practical and scientific significance of this work is the following: for a particular type of piston automobile pneumatic engine with a spool-type air distribution the level and patterns of changes in the energy loss of the operating fluid in the form of the enthalpy exergy loss in the intake system depending on the compressed air supplied pressure at the intake in the most important operating range of ps = 0.5-1.1 MPa at constant air temperature equal to ambient temperature Ts = Toc= 293 K have been determined.
The level of energy loss of the operating fluid (compressed air) in the form of its enthalpy exergy loss in the intake system of a four-cylinder piston pneumatic engine while its operating with various intake pressures are: with intake pressure ps = 0.5 MPa is 13.3 kW (116.5 % of Es), with intake pressure ps = = 0.7 MPa is 15.0 kW (135.3 % of Es), with intake pressure ps = 0.9 MPa is 16.9 kW (150.7 % of Es), with intake pressure Ps= 1.1 MPa is 19.7 kW (155.6 % of Es).
References
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References
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Voronkov Oleksandr Ivanovych, D.Sci., Associate
Professor of Internal Combustion Engines
Department, +380505830045, rio@khadi.kharkov.ua
Charchenko Anatolij Ivanovych, Associate
Professor of Internal Combustion Engines
Department, +38093549435, rio@khadi.kharkov.ua
Nikitchenko Igor Mykolayovich, PhD., Associate
Professor of Internal Combustion Engines
Department, +380993116110, nik@khadi.kharkov.ua
Novikova Yevgenia Borysivna, PhD., Associate Professor of Foreign Languages Department, +380958297372, rio@khadi.kharkov.ua
Teslenko Eduard Viktorovich, Post-Graduate of Internal Combustion Engines Department, +380678538385, rio@khadi.kharkov.ua
Nazarov Artem Oleksandrovych, Post-Graduate of Internal Combustion Engines Department, +380996537401, rio@khadi.kharkov.ua Kharkiv National Automobile and Highway University, 61002, Ukraine, Kharkiv, Yaroslav Mudryi st., 25.
Розрахунково-експериментальне визначення втрат енерги робочого тша в системi впуску автомобильного поршневого пневмодвигуна з використанням ексергетичного методу
Анотаця Розглянуто методику застосуван-ня эксергетичного методу термодинамгчного анализу для визначення втрат енергИ робочого тша в системi впуску автомобшьного поршневого пневмодвигуна з використанням експеримен-тальних даних його стендових моторних випро-бувань. Викладено анализ виконаного ексер-гетичного розрахунково-експериментального до^дження з оцтки величин енерговтрат у си-стемi впуску двигуна для актуальних режимiв його роботи.
Створений на кафедрi ДВЗ ХНАДУ для комбi-новано'1 автомобшьно'1 силово'1 установки зразок поршневого пневмодвигуна мае потребу в удоско-налюваннi й доводу, особливо вузлiв i систем, пов 'язаних зг здшсненням робочого процесу. Одтею з таких систем е система впуску стисне-ного повiтря. Для пневмодвигуна стиснене повiт-ря е не тшьки робочим тшом, але й енергоносiем для його системи впуску. До показниюв працез-датностi й ефективностi, ^iм надiйностi й дов-говiчностi, вiдноситься й показник рiвня енерге-тичних втрат.
Об'ектом до^дження е розглянутий пнев-модвигун, створений на кафедрi ДВЗ ХНАДУ шляхом конвертацИ бензинового чотирицилтд-рового чотиритактного ДВЗ.
Експериментальне до^дження проводилося шляхом лабораторних стендових моторних ви-пробувань пневмодвигуна зi зняттям швидюсних характеристик, за умови пiдтримки незмтних для кожноi характеристики двох термодинамiч-них параметрiв стисненого повiтря на входi: тиску ps= idem i температури Ts= 293 К. На кожному режимi випробувань рееструвалися ва зовнiшнi параметри роботи пневмодвигуна й mi-малися iндикаторнi дiаграми першого цилтдра. Кожна швидюсна характеристика складалася iз шести-восьми режимiв ps= idem i Ts= 293 К = = idem зi змiною частоти обертання колтчасто-го вала n, об/хв, вiд мiнiмально стшких (близько
n=200 об/хв) до максимально можливих n = 1000 ± 50 об/хв.
Сл1д зазначити, що в порожниш системи впуску вгдбуваеться падгння тиску робочого тыа внасл1док незворотних процесгв: дроселювання в контрольному отвор1, ггдравлгчних втрат по до-вжинг впускного каналу внаслгдок тертя газу об сттки, завихрень та iнших газодинамгчних явищ. Втрати ексергИ ентальпИ робочого тыа в сис-темi впуску DBrl = Евп-Енап становлять для тиску на впуску рх=0,5 МПа -13,3 кВт, для тиску на впуску = 0,7 МПа - 15,0 кВт для тиску на впуску = 0,9 МПа - 16,9 кВт , для тиску на впуску рs = 1,1 МПа - 19,7 кВт.
Ключовi слова: поршневий пневмодвигун, система впуску, енергетичнi втрати, ексергiя ентальпИ, режими максимальноi ефективно'1 по-тужностi.
О. I. Воронков, проф., д.т.н., А.1. Харченко, доц., к.т.н., I. М. Шмтченко, доц., к.т.н., С. Б. Новикова, доц., к.ф.н., Е. В. Тесленко, асшрант, А. О. Назаров, астрант, ХНАДУ
Расчетно-экспериментальное определение потерь энергии рабочего тела в системе впуска автомобильного поршневого пневмодвигателя
с использованием эксергетического метода
Аннотация. Рассмотрена методика приложения эксергетического метода термодинамического анализа для определения потерь энергии рабочего тела в системе впуска автомобильного поршневого пневмодвигателя с использованием экспериментальных данных его стендовых моторных испытаний. Изложен анализ выполненного эксергетического расчетно-эксперименталь-ного исследования по оценке величин энергопотерь в системе впуска двигателя для актуальных режимов его работы.
Созданный на кафедре ДВС ХНАДУ для комбинированной автомобильной силовой установки образец поршневого пневмодвигателя нуждается в совершенствовании и доводке, особенно узлов и систем, связанных с осуществлением рабочего процесса. Одной из таких систем является система впуска сжатого воздуха. Для пневмодига-теля сжатый воздух является не только рабочим телом, но и энергоносителем для его
системы впуска. К показателям работоспособности и эффективности, кроме надежности и долговечности, относится и показатель уровня энергетических потерь.
Объектом исследования является рассматриваемый пневмодвигатель, созданный на кафедре ДВС ХНАДУ путем конвертации бензинового четырехцилиндрового четырехтактного ДВС.
Экспериментальное исследование проводилось путем лабораторных стендовых моторных испытаний пневмодвигателя со снятием скоростных характеристик при условии поддержания неизменных для каждой характеристики двух термодинамических параметров сжатого воздуха на входе: давления ps= idem и температуры Ts= 293 К. На каждом режиме испытаний регистрировались все внешние параметры работы пневмодвигателя и снимались индикаторные диаграммы первого цилиндра. Каждая скоростная характеристика состояла из шести-восьми режимов ps= idem и Ts= 293 К = idem с изменением частоты вращения коленчатого вала n, об/мин, от минимально устойчивых (около n=200 об/мин) до максимально возможных n = 1000 ± 50 об/мин.
Следует отметить, что в полости системы впуска происходит падение давления рабочего тела вследствие имеющих место необратимо-стей: дросселирования в контрольном отверстии, гидравлических потерь по длине впускного канала в результате трения газа о стенки, завихрений и прочих газодинамических явлений. Потери эксергии энтальпии рабочего тела в системе впуска D^E^-E^ составляют для давления на впуске р=0,5 МПа - 13,3 кВт, для давления на впуске р=0,7 МПа - 15,0 кВт, для давления на впуске р=0,9 МПа - 16,9 кВт, для давления на впуске р=1,1 МПа - 19,7 кВт.
Ключевые слова: поршневой пневмодвигатель, система впуска, энергетические потери, эксергия энтальпии, режимы максимальной эффективной мощности.
А. И. Воронков, проф., д.т.н., А.И. Харченко, доц., к.т.н., И. Н. Никитченко, доц., к.т.н., Е. Б. Новикова, доц., к.ф.н., Э. В. Тесленко, аспирант, А. А. Назаров, аспирант, ХНАДУ
