Fren donanım kompresörlerinin performansı

Abstract

Bu çalışmanın esas temelini kamyonlarda kullanılan fren donanım kompresörleri oluşturmaktadır. Başlangıçta kompresörler sınıflandırılmış ve pistonlu kompresörlerin termodinamik analizi yapılmıştır. Daha sonra fren donanım kompresörlerinde dizayn ve uygulamada bilinmesi gereken önemli hususlar verilmiştir. Ayrıca pistonlu kompresör lerin en önemli sorunlarından biri olan çıkış havasına yağın karışmasına ve kompresör yağ kaybına sebep olan etkenler belirtilmiştir. Temel konular ele alındıktan sonra piston segman yuvalarında meydana gelen hava kaçaklarının (blow-by) teorik incelemesi yapılmış yarık akış modeli açıklanmıştır. Daha sonra ARMAS - İTÜ-KOSGEB işbirliği ile yeni geliştirilen A300 fren donanımı pistonlu kompresörünün deney standının kurulmasına ve bu kompresörün performans çalışmalarına yer verilmiştir. A3 00 kompresörü bu çalışma kapsamında kurulan deney tesisatında çeşitli testlere tabi tutulmuş ve deney sonuçlan teorik çalışma ile karşdaşünlrnıştır. Ayrıca bu kompresör aynı strok hacminde benzer bir kompresör ile kıyaslandığında daha üstün performans gösterdiği görülmüştür.

In the process industry, numerous systems operated with pressured gas is commonly designed because pressured gas has a working ability. Compressors are used to rise the pressure of a compressible fluid dependant on the design situations. Compressors has a numerous types and can be either with intermittent flow or continuous flow. Intermittent type compressors is named the positive displacement compressors. It can be seen compressor types following. ÇCOMPRESSORS > Continuous flow comp. Ejector comp. »MMWIMM MUUUUUUWUWUWUUUWUUWWUM Dynamic comp UUWMMmWWWW (Centrifugal,, Intermittent flow comp. Figure 1. General compressor types Reciprocating compressors which used in brake systems compress gas by reducing a volume of air isolated in the cylinder. The work per cycle or pressure- volume cycle can be seen and measured by indicator charts. The gas to be compressed all continues following operation. 1 -First, gas is isolated within a volume. 2-Then it is compressed to discharge pressure 3-Consequently, released into the discharge systems.Main characteristics of reciprocating brake compressors: Volumetric efficiency, T|v, in real is expressed to actual volume discharged per cycle over displacement volume per cycle and written as following equations: vp where Vs strok volume, Vp the volume actually pumped, T|v volumetrik efficiency. Compressor throughput per hours Vth (m3/h) is calculated by, Vth = 60.L.N.S where S cross-sectional area of the piston (m2), N shaft speed (rev./min.), L piston strok length (m). Real compressor throughput per hours V (m3/h) is calculated by, V=T|vo-Vth where V^ compressor throughput per hours (m3/h), tjvo total volumetric efficiency. Mean piston speed Cm (m/s) is defined as Cm=L-N/30 where L piston strok (m), N shaft speed (rev./min.). Discharge temperature T2; T2=T,(P2/P,)(n-l)/n where T suction temperature (K), P2,P, succesively effective exhaust and suction pressures (bar), n polytropic exponent of compression. The power consumptions of intermittent compressors can be obtained with following equations. XI^=0.028-V-Zave "lınek ' Mady n-1 where W compression power (kW), V total compressor throughput (m3/h), Zave average compressibility factor (=(Z, +Z2)/2), r|mek compressor mechanical efficiency at nominal speed (= 0.95), X]^ adiabatic efficiency of the cylinder(=0.88) ve n polytropic exponent of compression. Cylinder clearance c0: It is left a small clearance between the cylinder head and the piston to provide for thermal expansion and for machining tolerances. This clearance c" "0> co=0.005L+0.5 [mm] where L is the strok of piston (mm). The model for compressor crevices: This crevice-flow model is illustrated for air leakages between cylinder-ring- piston crevices. It is really an important problem to leak air from cylinder to crankcase. This leakages increases while piston travelled to the upper end of its stroke and rising upstream pressures. Moreover, after oil passes between cylinder- piston-ring, it mixes to the air. So it can cause fittings to become dirty. Assumptions: 1-It is accepted to have a uniform pressure throught the region. 2-In all regions, gas consumption is equivalent each other. 3-At crevice gas throughput is laminer 4-It is assumed that the pressure in the region 1 to be the same as changing cylinder pressures. 5-It is assumed that the pressure in the region 5 to be the same as crankcase pressure. 6-The piston is centered in the cylinder bore. 7-Cylinder temperature is assumed to same as the piston temperature. This crevice flow model is used to define flow between the piston, cylinder and the ring crevices. At this section, first for the regions the continuity equations below are written after making assumptions. dm ~d7 = Imç-Zmg XllFor the region 2 the region 3 the region 4. _ V2 dP2 mi2"m23"RT'"dT V3 dP3 mo + m23 _ m34 _ m35 = rj ' ~dT V4 dP4 m34"m45 = RT""dT is formulated by the continuity equations. Here Pj(i=l,2,3,4) the pressure of the region i, mij mass flow rate through either a ring gap or ring-side clearance. The mass-flow rate through the ring gap is defined from the orifice equation. m = pcCdAgZj where Cd discharge coefficient (=0.86), Ag ring end-gap area (m2), p upstream gas dencity (kg/m3), c speed of sound (m/s), Z{ compressiblity factor for isentropic flow. Through the ring-side clearance to determine the mass-flow rate, the flow is treated as an isothermal compressible flow through a narrow channel of height h and length Wr (.\2 m A V J h2(pu2-p/) 24W,H6" RT, where D=2h is the hydraulic diameter, A is the area normal to the flow, Ts is the ring temperature, R is the gas constant, Wr is the ring width, Pu and Pd is the pressures upstream and downstream respectively, \xgas is the viscosity of the gas. Experimental Studies: About A300 reciprocating compressor: The compressor A300 is used for air brake systems on trucks. It is developed with ARMAS-I.T.Ü. KOSGEB cooperation for getting rising trucks tonnages. XlllBasic compressor specifications: Type:A300 Rated:300 cm3 Bore: 92 mm Speed:3000 r/min. (max.) Operating pressure: 6-8 bar Suction pressure:Absulute 1 bar Oil temperature at gallery:40 °C Water flow:~7 m3/min. Strok length: 45. 1 mm Experimental Aparatus: It was necessary to make experimental aparatus to research the prototipe compressor A3 00 and the original HOLSET and to do performance analysis. For this reason experimental sets was established the base of this studies. The aparatus used for brake compressors consisted of experimental table and electrical motors, accupling units on it. At the same time the aparatus concluded inlet air filter and unloader valf controlling the receiver pressure. Cooling of the compressor was achieved by water circulation on its head. Therefore there were water pump and box at the aparatus. Also oilling of the compressor was achieved by oil circulation and controlling the oil pressure 4 bar at the system. After accupling the compressor, starting the unity and working until steady state, the temperature of the air at the compressor suction and exhaust was determined with termocuple resistance thermometer. The pressure of the air at the compressor exhaust was meassured by means of manometer. In practise, performed validation tests: 1- Durability test 2- Leakage test 3- Power absorbtion test 4- Pressure build up times 5- Oil carry over test After fulfilling performance tests and taking datas, it is determined diagrams for the compressors the A3 00 and the HOLSET. In this studies, the compressor A3 00 after developing prototipe and setting up experiment equipment is tried to make performance tests and compared with original compressor (HOLSET) having the same stroke volume. Consequently, measured datas has showed in the experiments that A300 brake compressor is well-quality and has a high performance.