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小型柴油機高速化和大功率化的發(fā)展 D．Broome Ricardo & Co.Engineers(1927) Ltd.(England) 筆者所在公司長期關(guān)注小型高速柴油機發(fā)展，特別是其燃燒系統(tǒng)的發(fā)展。僅管這類機型在北美大陸并不普及，但在歐洲和日本的產(chǎn)量以及使用量卻相當大，約有數(shù)百萬臺。典型的機型是自然吸氣四缸柴油機，每缸排量為25-35 in3 (400-600cm3)，工作轉(zhuǎn)速為4000-5000 rpm，活塞限速在2400ft/min(12m/s)。 在warren的美國軍方坦克研發(fā)中心（USATAC），Mich提議開發(fā)一種符合軍方要求的大功率輕型動力裝置，并且要求能達到前所未有的轉(zhuǎn)速以實現(xiàn)性能的提升，而不是僅僅增加渦輪增壓器。在這個情況下，進氣系統(tǒng)和燃燒系統(tǒng)的研究就被搬上了前臺。這項計劃制定了工作的執(zhí)行方案，包括設計，制造和在特定單缸機上的初步實驗。 項目 這個項目的技術(shù)要求最后由USATAC制定，其在大概的內(nèi)容如下所述： 1． 設計，采辦，制造和測試一臺缸徑為3.5英寸（88.9mm）的單缸機。要求以最高轉(zhuǎn)速工作，至少5000 rpm。以分離的空氣供給系統(tǒng)的對渦輪增壓狀況進行模擬。 2． 改進這臺單缸試驗機，以實現(xiàn)其預定的性能指標，使得相對應的軍用四缸機能夠產(chǎn)生 1 bhp/in3 (45.5kW/cm3)的功率以實現(xiàn)單位質(zhì)量功率3.5 lb/bhp (2.13kg/kW)。 3． 設計將不受傳統(tǒng)觀念影響，以最小化機械負荷和熱負荷為目標。 4． 最初的燃料要求以CITE-R fuel (MIL-F-46005A (MR))為標準。一開始，專門針對航空汽油進行研究，但后來這項要求有所放寬。 5． 如果可能的話使用MIL-L-2104B規(guī)格的潤滑油。 6． 項目的最后階段包括一臺軍用四缸機的設計，其中包含了對單缸機的測試。 7． 對于多缸機的啟動，怠速和小負荷工況的工作性能必須不能被忽視。 初步設計方案 一個對缸徑尺寸和功率輸出指標的試驗可以很快地顯示出引擎最高轉(zhuǎn)速的極限。從最小轉(zhuǎn)速5000 rpm開始，曲線清楚地顯示出轉(zhuǎn)速在6000 及以上時，這些活塞的運動速度與賽車引擎的相當。通過增加轉(zhuǎn)速來減小平均有效壓力的意義是重大的，fmep（這些估計都是根據(jù)筆者公司的一些過去的數(shù)據(jù)的分析，其中大部分被概括在圖1）的增加對指示平均有效壓力的減小幾乎沒有反饋。幾年來，自然吸氣柴油機被要求嚴格工作在冒煙界限以內(nèi)，因此其指示平均有效壓力只能提高到145lb/in2 (1000kPa)；因此，渦輪增壓的測量方面有了一些要求。高速和高引擎磨擦帶來的惡果就是燃料的消耗。表格1中清楚地顯示了bsfc對曲線的迅速惡化，效果并不比汽油機來的好，所以會損失壓縮做功循環(huán)的主要利益。根據(jù)這種情況，就計劃將轉(zhuǎn)速限定在6000 rpm。 主要的表現(xiàn)問題在引擎的設計中所要求的可以被概括為以下幾項: 引擎的進排氣—— 根據(jù)以往在高速小型柴油機上的經(jīng)驗表明，進排氣性能是活塞高速化時限制指示平均有效壓力的重要因素。因此，需要提供足夠大的充氣系數(shù)使得活塞速度能達到3500 ft/min (17.8m/s)，比現(xiàn)有的柴油機水平高出50%。這就必然要求設計脫離傳統(tǒng)的氣缸設計，包括更傾向于多氣門的設計（表2）；渦輪增壓帶來微小的變化，比如更高的進氣溫度，最小的壓力損失和減小體積功率的變化。 除此之外，渦輪增壓匹配要求需要被事先確定。對于車用引擎，由于對后備扭矩的需要，通常傾向于在額定的轉(zhuǎn)速前提下，最小化有效的推進力。因此大尺寸的排氣門并不被強求。另外，絕對短的廢氣排放系統(tǒng)要求排氣速度盡可能低，排氣門面積與進氣門相當。 同時需要考慮的是氣門開啟時間的設置。在進氣口，高速度通常聯(lián)系著較大的氣門關(guān)閉延時角，在柴油機里，一般為提前下止點45度，這將會連續(xù)地影響起動性能，同時也犧牲一定的低速性能，更大程度上影響到引擎的自然扭矩儲備。對于排氣，因為渦輪增壓的需要，排氣門在做功沖程結(jié)束前就要開啟，在高轉(zhuǎn)速下提前60度開啟并沒有什么優(yōu)勢。較大的氣門重疊角可以減少廢氣量和降低排氣系統(tǒng)各組成部分的溫度，同時可以最小化高轉(zhuǎn)速時充氣系數(shù)的降低程度。充氣系數(shù)與氣體的流動有關(guān)。而由于活塞和氣門間余隙容積產(chǎn)生的機械問題將會對燃燒系統(tǒng)有負面的影響。 燃燒問題——當引擎達到預定轉(zhuǎn)速時，點火延時期的持續(xù)時間將是一個基本的問題。點火延時持續(xù)時間將是在引擎正常工作時影響轉(zhuǎn)速，壓縮條件和燃油噴射提前角的重要因素，而特殊的燃燒室結(jié)構(gòu)將會是另一個影響因素。CITE-R燃油的最小十六烷值為37，但其已測定的數(shù)據(jù)只適用在低速狀態(tài)下，并不能直接應用在6000 rpm 的高轉(zhuǎn)速引擎的預測中。然而，可以從這些有用的數(shù)據(jù)中，得到小型高速引擎在使用輕油（十六烷值為55）時，以最小可起動壓縮比時的表現(xiàn)，估測如表1所示（更高的壓縮比會導致額外的熱損失和最大氣缸壓力的增長）。這些數(shù)據(jù)指出在6000 rpm時使用CITE燃油是可行的。但是在低負荷的情況下，要求入口溫度保持在室溫下，這對于采用渦輪增壓的機型是不可能的。 表 1 – 達到要求轉(zhuǎn)速時發(fā)動機的速度特性* 發(fā)動機轉(zhuǎn)速 5000 6000 7000 平均活塞速度, ft/min (m/s) 過 2915 3500 4085 (1480) (1775) (2070) 要求的bmep 轉(zhuǎn)速, lb/in2 (kPa) 158 132 113 (1090) (910) (780) 預測多缸機的 fmep, lb/in2 (kPa)** 72 83 95 (495) (575) (655) 所有發(fā)動機的Imep, lb/in2 (kPa) 230 215 208 (1585) (1485) (1435) 多缸機的Bsfc , 采用0.33 lb/inp.h(0.20 kg/kW.h), 0.48 0.54 0.61 Lb/inp.h (lb/bhp.h (kg/kW.h) (0.29) (0.33) (0.37) * 2-1/2×3-1/2 in (88.9×88.9 mm) 缸徑的發(fā)動機在 1 bhp/in3 (45.5 kW/cm3) 功率輸出。 ** 見圖2 燃油噴射時間由燃油噴射系統(tǒng)依靠后置式裝置控制，但是在現(xiàn)實中，開發(fā)新的噴射系統(tǒng)以替代傳統(tǒng)的脈沖式噴油泵系統(tǒng)是不可能的。在一個固定的節(jié)流孔噴口，需要考慮的是如何在6000 rpm 下，以全負荷運轉(zhuǎn)時能有60mm3 的噴射量，這就要求能提供合適的特性曲線，調(diào)整壓縮比在11：1。直噴式的燃燒系統(tǒng)在高轉(zhuǎn)速下延長噴射時間，其結(jié)果是很嚴重，這就不得不采用固定的節(jié)流孔噴口。另外一個附加的問題就是大負荷下DI系統(tǒng)工作的困難性，特別是在更高的冒煙界線空燃比情況下，要求更大的功率以達到預定的指標。在Ricardo的早期研究中已經(jīng)表明，基于慧星漩渦式的燃燒室系統(tǒng)（見表3），工作在4500 rpm 的小型高速民用引擎通過改進可以達到特殊應用的目的。僅管有獨立的起動裝置和完全多油路系統(tǒng)，但是主要可預見的問題是在高熱負荷的情況下，并沒有合適的DI系統(tǒng)；盡管如此，隨著可應用于多缸機的內(nèi)置式輔助裝置的應用，以及CITE燃油的限制，這些問題都不會很嚴重。 表 2 – 發(fā)動機進排氣系統(tǒng) - 氣門面積和限制速度* 合適的氣缸蓋布置的形式 氣門面積比例 相對于缸徑的百分比 結(jié)果 進氣 排氣 Total rpm,max 傳統(tǒng)的兩氣門自然進氣 18 12 30 3900 平頂燃燒室四氣門渦輪增壓 17 17 34 4000 斜屋頂三氣門渦輪增壓 21 21 42 5000 斜屋頂四氣門渦輪增壓 25 25 50 6000 * 3-1/2 in (88.9 mm) 發(fā)動機行程. 受限制的平均進氣氣體速度: 自然進氣,210 ft/s (64 m/s); 渦輪增壓, 230 ft/s (70 m/s). 在項目開始的時候，一些DI系統(tǒng)被拿來做選擇，除此之外有一臺4000 rpm的試驗用單缸機以及相配套的已設計完成的DI系統(tǒng)版本。在那段時期，受制于噪聲，煙氣和獨特的廢氣排放法規(guī)，分離式燃燒室受到更多的重視，而在DI系統(tǒng)上的試驗工作并沒有像現(xiàn)在一樣被提上議程。 引擎的磨擦問題——與同樣尺寸大小的常規(guī)民用多缸高速渦輪增壓柴油機相比，很明顯，通過提高轉(zhuǎn)速和活塞速度可以顯著提高fmep。在表1中可以清楚地看到，如果想到同時達到功率輸出和燃油消耗率兩項指標是一個很大的問題。圖2所示是一臺典型的民用引擎fmep/速度曲線圖，其估測的fmep是從高速多缸機的試驗中獲得的。在這個測試中，一些柴油機機構(gòu)基本的機械磨擦的增值已被假定，直到渦輪增壓裝置增加氣缸壓力和采用更大的軸承以獲得可接受的可靠性。另外，進排氣的泵氣損失將會大幅增加高速機的?fmep，除非能夠設計一個合適的氣門直徑并長期保持不變。 表3 – 燃燒室特性 慧星式 直噴式 漩渦式燃燒室 燃燒室 特殊轉(zhuǎn)速 (受限于 A/F) 好 失敗 燃燒控制和機械負荷 好 差 熱負荷和熱損失 差Poor 好 獨立起動性 差Poor 好 采用加熱塞的起動性能 好Good ---- 軸針類型 針式 多孔式 排放 (NOx) 好Good 差r 多燃料性能 差 失敗 按照汽油標準，該機型的單位機械效率將會很低，但經(jīng)驗告訴我們，盡管在細節(jié)方面的設計可以獲得額外的收益，但是較低的水平仍然會隱藏在設計的規(guī)格里。 單缸試驗機 基于對外形的考慮，單缸試驗機的最后設計被確定，其基本尺寸為：缸徑×沖程： 3-1/2 in ф×3-1/2 in (88.9mmф×88.9mm)；正常滿負荷轉(zhuǎn)速范圍，3000-6000 rpm；最高氣缸爆發(fā)壓力，2500 lb/in2(17.3Mpa)。 氣缸爆發(fā)壓力也許比傳統(tǒng)標準定得稍高，過去的經(jīng)驗顯示，設計如此一臺引擎是很危險的，所以在試驗過程中會造成不可遇見的局限性。事實上，最初可正常工作的DI版本的設計為3000 lb/in2(20.7Mpa)，但在隨后的慧星版本中有一定的減小。 慧星渦旋燃燒室引擎的布局如圖3-5所示，整機的主要部件如圖6所示，如下所述： 曲軸箱——曲軸箱及其后蓋由球墨鑄鐵（BS 1452：1961，Grade 14）鑄成，兩者通過螺栓聯(lián)接。曲軸箱延用Ricardo設計的E/6 可變壓縮比汽油機的曲軸箱，這就導致了燃燒室如E/6型機那樣，置于曲軸箱前端。有三個主要軸承，全部采用鉛青銅合金襯套，中間一個軸承采用止推軸承。后軸承作為曲軸的延長軸的固定軸承，可以進行調(diào)節(jié)軸向位置，所以不能分擔中間軸承承受的燃燒負荷。 曲軸——曲軸由一次鍛造成形的滲氮鋼，BS 970:1955 En 40c。平衡重為整體式，只用來平衡旋轉(zhuǎn)慣性力，同時設有平衡一階和二階往復慣性力的平衡軸。所有主軸頸和曲柄銷表面滲氮處理，三處主軸頸直徑，從前到后，分別為3,3,and 2-3/8 in (76.2,76.2,and 60.4 mm)，曲柄銷直徑為2-5/8 in (66.6mm)。 連桿——為獲得較好的模具和避免多余的費用開支，連桿的方案從民用機型上進行選擇，最后選定了福特 2700 系列柴油機的連桿。 連桿大頭可承受的最大氣缸爆發(fā)壓力為3000 lb/in2，連桿小頭則通常不被考慮。因此在將該型連桿運用到慧星漩渦燃燒室版本的引擎上時，要求其最大氣缸爆發(fā)壓力為 2500 lb/in2。軸承采用2700系列引擎上的型號，由15%的錫鋁合金做成襯套，而小頭襯套采用預制的鋁青銅合金。除了在加工時要求仔細拋光和檢驗連桿質(zhì)量外，連桿小頭應盡可能減小寬度，以減小活塞銷座和活塞銷襯套的熱負荷和慣性負荷。加在連桿大頭固定螺栓上的扭矩高于傳統(tǒng)的標準，以防止在6000 rpm時，當活塞到達排氣上止點時螺栓帽脫落。 軸承的提供者The Glacier Metal Co.Ltd. 出了計算機計算的結(jié)果，表明所提供的軸承的工作范圍是合適的，盡管當高速時連桿大頭將承受相當大的慣性力。 活塞和活塞銷——活塞由含硅13%的鋁合金整體鑄造而成，性能為BS 1490:1970 LM13WP，在斜屋頂燃燒室的有角度的一面上開有淺槽和雙凹槽?；钊捎脙傻罋猸h(huán)，第一道為桶形環(huán)，第二道為錐形的扭曲環(huán)；油環(huán)也作相應的設置。為提高耐磨性，在環(huán)的磨擦面上鍍銅。不鍍鉻是因為鍍銅的環(huán)性能已經(jīng)足夠了。 雖然活塞高度太大（相對于柴油機的普遍標準）會使設計合適的活塞，活塞環(huán)和缸套變得困難，但是在這個機型里活塞仍按照實際需要被特意設計成較高的高度。盡管如此，這一些小問題根據(jù)被經(jīng)驗所克服。 活塞的冷卻和連桿小頭的潤滑是通過一個安裝在曲軸箱內(nèi)的固定噴嘴定時噴射實現(xiàn)的。這個方法可以取消連桿大頭軸瓦處油槽的設計，從而使軸瓦完整。兩個活塞的改進，盤式冷卻系統(tǒng)的安排，可形變式核心的設計都如圖所示。為獲得可行的冷卻方案，對活塞環(huán)帶的盤式冷卻系統(tǒng)進行設計，承受從活塞頭部傳過來的氣體壓力的支桿被設計得近可能細，在這個區(qū)域里允許有額外的扭曲。形變式核心有著卓越的性能。 活塞銷材料為淬火鋼，直徑為1-3/8 in (34.9mm)： 缸套和水套—因為使用漩渦式燃燒室系統(tǒng)，在局部有較高的熱傳遞速率，伴隨有 2500 lb/in2的氣缸最大爆發(fā)壓力，使得在設計濕缸套的時候有相當?shù)碾y度。根據(jù)傳統(tǒng)的鐵制氣缸套厚度設計，則會使第一道環(huán)反向點的表面溫度額外地上升。最后方案選擇使用鋼制氣缸套，在內(nèi)壁鍍上一層厚度為0.0015 in (38μm)的堅硬的鉻合金。越向頂部，氣缸套越薄，使得溫度能得以控制，但是較小的厚度將會削弱缸套的剛度，不利于抵御有水一側(cè)的沖擊。缸套的頂端被折邊以適當?shù)倪^盈配合安裝在氣缸上部，下端裝有水封安置在曲軸箱頂部；其徑向定位由曲軸箱內(nèi)的氣缸螺栓提供，水封為通常的橡膠環(huán)。 氣缸蓋總成——氣缸蓋以及連接架，是發(fā)動機最復雜的部件，現(xiàn)在依舊有不少設計問題有待解決。如何設計合適的氣門和氣道成為其具有相當難度的核心問題。在傳統(tǒng)的缸徑為5-1/2 in的柴油機中為氣門留有3-1/2 in的直徑位置，這樣可以為漩渦式燃燒室系統(tǒng)的活塞區(qū)域提供高達5.7 ihp/in2 (0.66 indicated kW/cm2)的冷卻能力。 僅管四氣門的布置在其它地方得到了充分的驗證，但是在這個慧星漩渦式燃燒室做四氣門的布置并不十分恰當，因此最后選定了三氣門的布置方案。斜屋頂?shù)念^部表面積是很重要的，一部分留給必要的氣門面積，但更多是用來預留額外的散熱空間。通常情況選擇使用兩個排氣門，以減小它們各自的尺寸，同時使用一個進氣門。但是最終一個相反的方案被選定，同圖8所示。一對進氣門布置在中央，因為考慮到如果采用一對排氣門的話，較長的排氣管會對燃燒室頂部強烈地加熱，對排氣門設計不利。不對稱的燃燒室和氣門布置以前也被研究過，但因其不能顯示出實際的優(yōu)勢和與多缸機的要求不符合而被舍棄。在這個設計中，通過將加熱塞置于燃燒室較低的位置，采用所謂的延展式外形的慧星燃燒室以最小化空間。 為了冷卻進氣門、排氣門以及兩者之間的橋梁結(jié)構(gòu)，鉆孔是不可避免的，這就需要對燃燒室與水套之間的金屬層厚度做精確的控制，并且要求水套的表面保持清潔。漩渦燃燒室的加熱塞置于窄處，由Nimonic 80A合金精密鑄造而成，是一個輕型的定位于銅襯墊的上邊緣口，根據(jù)經(jīng)驗這樣的強化在一些地方需要直接冷卻，而不像民用發(fā)動機那樣存在一個間隙用來迅速暖機。在燃燒室上部有一個由球墨鑄鐵制成的噴嘴。一個傳感器的觸點置于氣缸頂部的后端。 輕型合金機座直接安裝在缸套的上邊緣，沒有用到襯墊，通過8個單頭螺柱垂直夾緊在水套的上邊緣。根據(jù)經(jīng)驗，燃氣的沖擊不構(gòu)成問題，設計的單頭螺柱滿足1.4倍于滿負荷燃氣沖擊的要求。 進排氣的頂置凸輪軸采用11個螺栓，通過三角架布置在頂部，以確保安全。凸輪軸通過搖臂驅(qū)動氣門的開啟和關(guān)閉，氣門間隙的調(diào)節(jié)通過對氣門挺柱頂端的墊片來實現(xiàn)。凸輪軸對于進氣門關(guān)閉和排氣打開的時刻的控制是可行的。進氣門的材料采用鋼BS 970:1955 En 59 “XB”，對于單獨的排氣門采用特殊的21-4/n奧氏體鋼，但根據(jù)經(jīng)驗，排氣門需要承受高溫，在頭部會發(fā)生穴蝕現(xiàn)象，因此在頭部采用Nimonic 80合金，并鑲有鉻合金。 為了實現(xiàn)多缸機在質(zhì)量上的指標，整機廣泛采用鋁合金材料，包括氣缸蓋和機體部分。氣缸蓋初始設計采用BS 1490:1970 LM25WP Al-Mg-Si合金，但該鑄件卻被證實在進氣口附近存在多孔滲水現(xiàn)象。因此在鑄造方面做了相關(guān)改進，材料也被改成高溫RR 350 Al-Cu-Ni-Co-Sb-Zr 合金。之后鑄件就再也沒有出現(xiàn)過大的松孔，但隨之而來的是進排氣門之間的橋梁結(jié)構(gòu)出現(xiàn)破裂現(xiàn)象。冶金實驗證明微孔收縮和薄膜都是在鑄造中形成的，但是這些缺陷是非常嚴重的，通過一對固定的熱電偶對稱置于橋梁結(jié)構(gòu)不同深度處，以選擇合適的插入點的方法，因為物理上的原因在這里是不適用的。如表4所示的這些實驗中，證實按照最初的設計程序計算得出來的數(shù)據(jù)，如圖4所示，采用Lm25WP 的合金時，材料處于正常工作范圍的邊緣，采用RR 350時材料失效并不僅僅是熱疲勞影響。在對設計做了一定的修改后，加強了橋梁結(jié)構(gòu)的冷卻和強度，同時減小了進排氣門的直徑，分別為1.050 in (26.7mm)和1.505 in (38.2mm)，在室溫下經(jīng)過機械負荷測試時，得知缸蓋甚至只是在裝配后就會發(fā)生一定的扭曲變形。微型應變儀被應用在這些橋梁結(jié)構(gòu)，氣門座，氣缸蓋總成，裝配，整機的扭矩和氣缸體所承受的氣缸爆發(fā)壓力是很合適的。測定的結(jié)果，與一臺相似的技術(shù)成熟的，采用鋁制氣缸蓋的民用兩氣門柴油機相比較，比較結(jié)果見圖9，證明在冷態(tài)時兩種機型的橋梁結(jié)構(gòu)都存在明顯的預應力，但是在發(fā)動機工作會被熱應力負荷所抵消。對RR 350材料的冷態(tài)結(jié)果分析表明，在溫度上升后，有足夠的安全系數(shù)保留，也沒有明顯的缺陷問題。 表4 – 氣缸頂部中央的熱流和溫度* 估測 標準 局部熱流, 518,000 475,000 Btu/ft2.h (MW/m2) (1.63) (1.5) 缸壁一側(cè)的溫度 504 525 F (C) (262) (279) 推斷 * 轉(zhuǎn)速: 215 lb/in2 (14.85 bar) at 6000 rpm. 現(xiàn)在設計的氣缸蓋，已經(jīng)證明其失效時間為早期設計的2倍，可以說以上所說的問題都已經(jīng)克服，一個可行的氣缸蓋設計已經(jīng)成形。 正時驅(qū)動平衡齒輪——一個后置蓋將整個齒輪傳動系統(tǒng)罩住，下方是兩個主平衡軸傳動齒輪和兩個二級兩倍速的平衡軸傳動齒輪，在上方是一個一倍的減速齒輪。平衡軸位于曲柄行程的下方。平衡軸以鋼棒材為毛坯，一端磨削后壓入薄鐵管，使其有光滑的外表面以減少動力損失和機油的攪動。從定時齒輪箱伸出的半速輸出軸前端與噴油泵相聯(lián)，后端與齒形帶輪聯(lián)接。 兩根頂置凸輪軸通過1：1的同步帶輪與位于缸蓋總成齒輪的副軸相聯(lián)接；這根副軸有一定的撓性，以允許將來需要改進時，燃燒室的改變和氣門位置的布局能夠得以實現(xiàn)。深溝球軸承被應用于正時傳動系統(tǒng)。 燃油噴射系統(tǒng)——在更早的時候，就開始采用傳統(tǒng)的高壓燃油噴射泵系統(tǒng)，之后又開始對適用在本機型，轉(zhuǎn)速在3000 rpm 的噴射泵的研究。該工作單元是從一臺適用于3000 rpm，兩沖程三缸機的噴油泵上改進而成的，其單獨的動力學原理需要在高速下得以驗證。為獲得足夠高的噴射速度和噴油量，一種新的凸輪軸被開發(fā)出來，這樣三個單獨的工作單元可以相互聯(lián)系起來。噴油泵通過一個手動的調(diào)速器控制，以使得在調(diào)節(jié)噴油時刻時不用停機。 在噴油泵附近靠近水套的地方允許有一短的高壓油管，而傳統(tǒng)的S號的軸針式噴嘴正好合適。民用發(fā)動機上的軸針熱屏蔽裝置可以用來保持軸針尖較低的溫度。 冷卻回路——通過一個外置的電機驅(qū)動冷卻液高速流動，同時在主循環(huán)上還安有一流量計。引擎的冷卻回路是一個平行系統(tǒng)，冷卻液同時從幾個入水口進入。在氣缸水套底部開有兩個互不干涉的孔，冷卻液通過這兩個孔經(jīng)過缸套和缸套邊緣的壓邊處。然后冷卻液經(jīng)12個在缸蓋和水套之間的缸蓋熱圈上的孔進入氣缸蓋。另一條水路是在氣缸頂部的橋梁結(jié)構(gòu)上鉆有四個孔，冷卻水流在進入主循環(huán)前先流經(jīng)進氣門周圍的橋梁結(jié)構(gòu)。缸蓋通過一個單獨的排水通道與排氣門相聯(lián)。 潤滑油路——同樣，一個單獨的機油泵用來提供潤滑油，以滿足對潤滑和活塞冷卻的要求；流量計安置在一條獨立的通路上。所有的潤滑油經(jīng)過機體內(nèi)分離的鉆孔和外部管道相聯(lián)，經(jīng)驗表明在典型的發(fā)動機上，這樣的布置可以最大程度上減少成本。低壓供油給氣門結(jié)構(gòu)。 機油箱是干式的，同時有一定的加壓以防止機油浸到平衡軸上。 發(fā)動機性能 對引擎的燃燒系統(tǒng)和噴射系統(tǒng)的改進并不是同時進行的，而且對發(fā)動機的進氣系統(tǒng)并沒有做最優(yōu)化設計。因此，這里所引證的數(shù)據(jù)只是表明指標要求能夠達到，而不是為了對發(fā)動機性能做最優(yōu)化設計。當然，更多的成果也是被期待的，特別是在超過轉(zhuǎn)速范圍時的平衡性和在超高速下燃燒系統(tǒng)的機械和熱負荷數(shù)據(jù)。 測試安裝——發(fā)動機與電子應變測量計相聯(lián)，也可與駕駛單元相聯(lián)接。冷卻液出口和潤滑油入口溫度可以通過水冷散熱交換器自動控制。壓縮空氣由獨立的空氣壓縮機提供，中間聯(lián)有一個ALCOCK滯流空氣流量計，另有一個中冷器，用來控制進氣溫度和減少沖擊效應。排氣管通過一根短管與一個膨脹室相連，在出口處壓力得到控制；在對發(fā)動機進排氣系統(tǒng)的增壓模擬中，一個氣室或一排管道被用來進氣。其測試安裝結(jié)果如圖10所示。 通常的測試方法是，通過改變噴油量，控制引擎工作在一定的負荷范圍以上，并保持一定的轉(zhuǎn)速，而進氣量和排氣管背壓保持與對照樣機相同。這樣并不是模擬增壓發(fā)動機工作在相同的負荷范圍內(nèi)，而是通過這些數(shù)據(jù)大致地預測增壓發(fā)動機的工作狀況。根據(jù)表現(xiàn)出來的發(fā)動機的特征結(jié)果，在相同的進氣量和背壓的情況下，對中斷結(jié)果和動力磨擦的情況的分析，允許對多缸機指標的推測。 由于CITE燃料在英國供給困難，USATAC贊成使用本質(zhì)上相同的航空用低烷值燃油（D.Eng R.D.2486），其十六烷值為37。一些實驗也使用C.I.油料，是一種本質(zhì)上與ASTM Grade 2-D柴油相似的燃油，其十六烷值為55。 測試工作——對這臺機械部分改進的新型的發(fā)動機，大部分測試工作表明其完全達到了215 lb/in2 (1485 kPa) imep的指標要求。一些數(shù)據(jù)也表明，在達到滿負荷工作運轉(zhuǎn)時，轉(zhuǎn)速會降到3000 rpm。 在開始測試時，在6000 rpm的轉(zhuǎn)速下其冒煙界線是可以接受的，但在排氣管邊緣的熱電偶測得的排氣溫度過高，達到了1650F (900C)。這是由于非常長的噴射時間和延遲燃燒造成的。對燃油噴射系統(tǒng)的改進減低了噴嘴的噴射速度，但是只是在名義上有大量的減少，在噴嘴口處，只有20%的減少量在噴射時間內(nèi)。對噴射系統(tǒng)的簡單計算出現(xiàn)另人失望的結(jié)果是自然的。因此就需要像一開始一樣，對高排氣溫度時固有的熱力問題在高排氣溫度進行研究和尋找設計解決方案。 在負荷范圍內(nèi)，在進氣壓縮比為1.9和6000 rpm時，兩個燃料的性能如圖11所示，而其典型的氣缸爆發(fā)壓力和軸針抬起高度如圖12所示。從以上數(shù)據(jù)可得知，增壓對于兩種燃料達到性能指標都是非常有效的，其在冒煙邊界的A/F值為0.055。這些值得注意的數(shù)據(jù)結(jié)果證明，最后燃油注入燃燒室的時刻為40 deg atdc，這可以歸功于慧星燃燒系統(tǒng)出色的混合氣能力，特別是在活塞凹陷處非常有效的氣體流動。在使用輕油時，慧星燃燒室可以達到最適宜的起動性能，但在使用對點火要求更低要求的航空低辛烷值燃油時，它就不可能達到在相同的進氣壓力下相同的燃燒效果，不可避免地會有些損失。在小負荷時不能發(fā)火的問題同樣也在使用航空低辛烷值燃油時體現(xiàn)出來，增加一定的進氣溫度可能成為必要的措施—多缸機使用一根進氣總管—采用對冷卻液特殊控制的二次冷卻器成為必要的手段。 從氣缸爆發(fā)壓力圖表上可得知，氣缸最大爆發(fā)壓力在高速的情況下，可以測得其值在滿負荷時為1475 lb/in2 (10.2MPa)。 在恒定的進氣壓力和最適宜的噴油時刻時，使用輕油的速度曲線如圖13所示，噴油時刻的改變?nèi)鐖D11所示。應該指出在燃燒室低轉(zhuǎn)速輸出時相對于高轉(zhuǎn)速下的要求有所降低，盡管隨著更大的改進也會有相應的改進?，F(xiàn)在相當小的氣門被應用，而早期的進氣門關(guān)閉時刻為40 deg abdc，其體積效率在5000 rpm時下降，但是多缸機對扭矩的儲備要求還是可以接受的。 多缸機的可能性 在對單缸機試驗完成時，還沒有認為值得花時間對多缸機的進行設計研究。但是，為了獲得對今后整機樣式的大致了解，一些可能的四缸機方案草圖還是有的，其前端視圖和側(cè)視圖如圖14所視。該發(fā)動機的正視圖表現(xiàn)了單列、干機油箱的設計，它的氣缸蓋除了采用整體式缸蓋外，本質(zhì)上與單缸機的相同。渦輪增壓器安裝在飛輪上方，與進入總管和中冷器相聯(lián)接。該引擎，包括各個輔配件，包括飛輪長31-3/4 in (806 mm)，高 32-3/4 in (831) mm)，寬24 in (609 mm)，單位質(zhì)量功率為9.3 bhp/ft3 (245 kW/m3)。為了達到特定的重量要求，其干重量必需達到470 lb (213 kg)。其質(zhì)量與排量比為3.49 lb/in3 (0.096 kg/cm3)，而傳統(tǒng)的車用柴油機的質(zhì)量與排量比為3-5 lb/in3 (0.08-0.14 kg/cm3)。 基于單缸試驗機的實驗數(shù)據(jù)，為該機型設計的制動器的性能也基本上達標了。在4000 rpm時要求有10%的扭矩儲備要求，因此，作為一臺渦輪增壓發(fā)動機，已經(jīng)達到了扭矩的要求。在4000 rpm以下，供油量需要削減以保持其排煙有約5%的不透明度，這樣在低速時就可能獲得一定的性能增長。進氣壓縮比被要求在通過中冷器損失后還有2.7，其最大的氣缸爆發(fā)壓力為1950 lb/in2 (13.5 MPa)。排放情況在單缸機上并沒有做測定，但是慧星漩渦式燃燒室已經(jīng)被證明其擁用極低的排放，特別是是氮氧化物方面，倘若不發(fā)火的問題能夠得到解決，那么這臺發(fā)動機滿足1975C.A.R.B的要求也就成為可能。 概括和總結(jié) 在USATAC發(fā)起的設計和測試過程中，達到高效的輸出1 bhp/in3 (45.5 kW/cm3)和低重量的柴油機的可能性也曾被研究過，采用高轉(zhuǎn)速和一般的活塞運動速度，分別為6000 rpm和3500 ft/min (17.75m/s)。預測要達到的目標是imeps達到 215 lb/in2 (1485 kPa)，因此也就需要采用渦輪增壓系統(tǒng)。 在研究一開始所涉及到的三個主要問題： 1. 引擎的進排氣系統(tǒng)，主要是盡可能提供比傳統(tǒng)發(fā)動機更大的氣門面積。 2. 燃燒，包括最基本的在很短的時間間隔內(nèi)實現(xiàn)壓燃的問題，也有燃油噴射系統(tǒng)和燃燒室的選型。 3. 發(fā)動機的磨擦，對bfc的控制。 從以上幾點考慮，一個單缸試驗機的設計提了出來，包括對進排氣系統(tǒng)和燃燒等細節(jié)問題的研究。 1. 使用Ricardo設計的慧星漩渦燃燒室系統(tǒng)。 2. 采用斜屋頂燃燒室頂部設計，以便提供足夠的氣門安裝空間。 3. 雙進氣門和單排氣門設計。 4. 合理地對機械部分做出設計，以應付在特殊地試驗條件下對很高氣體爆發(fā)壓力的抵抗要求，以及設計適應高熱應力的冷卻系統(tǒng)。 測試工作表明設計的指標可以達到，盡管改進工作還沒有完成，但是試驗數(shù)據(jù)顯示進排氣系統(tǒng)和采用慧星燃燒室的燃燒系統(tǒng)可以實現(xiàn)轉(zhuǎn)速6000 rpm，這還是在采用傳統(tǒng)的高壓燃油噴射泵的情況下。采用CITE燃油也可能是出現(xiàn)這種情況的原因。 最初為軍事目的設計的，背包大小，特殊功率要求的4缸試驗機已經(jīng)制造完成。 致謝 作者非常感謝文中所提及的美軍坦克研發(fā)中心在這個項目的大力支持，也很感謝他們和Ricardo & Co.對論文出版的許可。 參考文獻 1. B.W.Millington and E.R.Hartles, “Frictional Losses inDiesel Engine.” SAE Transactions, Vol.77 (1968), paper 680590. 2. C.J.Walder, “Some Problems Encountered in the Design and Development of High Speed Diesel Engines.” SAE Transactions, Vol.74(1966), paper 650025. 3. W.T.Lyn and E.Valdmanis, “The Effects of Physical Factors on Ignition Delay.” Paper 689192 presented at SAE Automotive Engineering Congress, Detroit, January 1968. 4. C.C.J.French, “Taking the heat off the Highly Boosted Diesel.” SAE Transactions, Vol.78 (1969), paper 690463. 5. C.J.Walder, “The Reduction of Emissions from Diesel Engines.” Paper 7320214 presented at SAE Automotive Engineering Congress, Detroit, January 1973. 21 Toward Higher Speeds and Outputs From the Small Diesel Engine D.Broome Ricardo & Co.Engineers(1927) Ltd.(England) THE AUTHOR’S company has long been concerned with the development of the small, high-speed diesel engine, and is particularly associated with combustion systems for this type of engine. Although such engines are not common in the North American continent, production and use in Europe and Japan is considerable, totaling several million units. These are, typically, naturally aspirated 4-cyl engines of 25-35 in3 (400-600cm3) displacement per cylinder, operating up to speeds of 4000-5000 rpm, with a limiting piston speed of about 2400ft/min(12m/s). In discussion with the U.S.Army Tank-Automotive Command (USATAC) at , Mich.it was proposed that the military requirement of high power from a small lightweight package could be achieved by exploiting higher speeds than hitherto, rather than the application of increased levels of turbocharger alone, and this led to the formulation of a research program to study combustion and breathing problems under such conditions. This paper describes the work carried out to date, which has involved the design, manufacture, and preliminary test work on special single cylinder engine. THE PROJECT The project specifications finally laid down by USATAC can be summarized as follows: 1. Design, procure, build, and test a single-cylinder engine of 3-1/2 in (88.9mm) bore and stroke, to operate at the highest possible speed, but certainly above 5000 rpm. Simulation of turbocharged conditions to be achieved using a separate air supply. 2. To develop the single-cylinder test engine to achieve performance targets such that a 4-cyl version for military duties could produce 1 bhp/in3 (45.5kW/cm3) displacement with a target dry weight of about 3.5 lb/bhp (2.13kg/kW). 3. The design not to be influenced by conventional practices, with the aim of minimizing mechanical and thermal stresses. 4. Operation on CITE-R fuel (MIL-F-46005A (MR)) to be the primary requirement. Initially, fuels down to aviation gasoline were to be investigated, but this latter requirement was subsequently relaxed. 5. Lubricating oils to the MIL-L-2104B specification to be used if at all possible. 6. The final phase of the project to include a design study for a 4-cyl military engine, embodying the lessons learned on the single-cylinder test unit. 7. Starting, idling, and light-load operation of the multicylinder engine must not be compromised. PRELIMINARY DESIGN CONSIDERATIONS A simple examination of the cylinder size and power output target rapidly showed the limitations that the maximum engine speed would have on performance (Table 1). Starting from the minimum speed specified of 5000 rpm, it is clear that speeds of 6000 rpm and above entail piston speeds equal to those of racing gasoline engines. While the reductions in bmep through use of high speeds are significant, the increases in fmep (estimated from past results obtained at the author’s company, much of which has been summarized in Ref.1) give very little return in reduced imep. Naturally aspirated automotive diesel engines working to the strict smoke limits of a few years hence can only operate up to about 145lb/in2 (1000kPa) imep; hence it was clear that some measure of turbocharging would be required. A further penalty of high speed and high engine friction is in fuel consumption, and Table 1 makes clear how the bsfc would worsen rapidly to levels no better than a gasoline engine, so losing one of the major advantages of the compression ignition cycle. In these circumstances,it was decided to limit the speed of the research engine to 6000 rpm. The major performance problems involved in the design of an engine to meet these requirements might be summarized as follows: ENGINE BREATHING-Previous experience on small high-speed diesels had shown that the major limitation on imep at high piston speeds is the breathing of the engine (2).Hence, valves of sufficient flow area had to be provided to allow efficient operation up to 3500 ft/min (17.8m/s) piston speed, some 50% higher than levels normally employed in diesel engines. This would certainly require departures from conventional cylinder head arrangements, involving inclined multiple-valve designs (Table 2);turbocharged operation brings a slight bonus in that the higher inlet air temperatures minimize pressure losses and reduce volumetric efficiency changes. In addition, possible turbocharger matching requirements had to be borne in mind. While, for automotive engines, torque backup requirements normally favor minimizing the available boost at the rated speed, so that a large exhaust valve area is not mandatory, in this case the very short absolute exhaust gas release periods suggested that the exhaust mean gas velocities should be kept low, and exhaust valve area about equal to that of the inlet. Also requiring consideration was the question of valve timings. For the inlet, high speeds are normally associated with a late closing point, yet in the case of a diesel, and with closing points later than about 45 deg abdc, there would be a progressive sacrifice in starting ability, as well as some loss of low-speed performance, which would further impair the natural torque backup characteristics of the engine. For the exhaust, the turbocharger matching requirement again dictates an early release of the gases on the expansion stroke, and timings later than about 60 deg abdc do not show to advantage at high speeds. While a long overlap period could contribute to reduction of exhaust gas and exhaust system component temperatures, such gains would be minimal at high speeds due to the very low quantity of scavenge air which might be passed relative to the trapped flow, and the mechanical problems of obtaining the piston/valve clearance would place a severe penalty on the combustion system. COMBUSTION PROBLEMS-A fundamental problem likely to affect the engine at the speeds contemplated was the likely duration of the ignition delay period. Ignition delay is a function of engine speed, compression conditions, and injection timing for a fuel of particular ignition delay is a function of engine speed, compression conditions, and injection timing for a fuel of particular ignition quality at normal running conditions (3),if factors related to the particular combustion chamber configuration in use are considered as being of second order. CITE-R fuel has a minimum specified cetane rating of 37,but the published data on engine delay using this fuel covered only low-speed conditions, and were not of direct use in predicting results at 6000 rpm. However, consideration of these available data, together with the known performance of small high-speed engines operating up to 5000 rpm on gas oil (55 cetane), led to the estimates shown in Fig.1,for the lowest compression ratio which would allow acceptable starting (higher ratios would give excessive heat losses and maximum cylinder pressures).These suggested that unaided or true compression ignition operation at 6000 rpm was feasible on CITE fuel, although the light-load condition would require inlet manifold air temperatures to be maintained significantly above ambient-not an impossible requirement for a turbocharged engine. Injection periods being controlled by the injection system will depend on the latter’s type, but for practical reasons there could be no possibility of developing new systems for the project, to replace the conventional jerk pump arrangement. With a fixed orifice area nozzle, there would be considerable problems in passing the required full load quantity of up to about 60 mm3/injection at 6000 rpm at the required rate, yet obtaining satisfactory characteristics for idling, the turndown ratio being about 11:1.The effect of an extended injection period on combustion at the high speeds required could be very severe on a direct injection (DI) combustion system, where use of a fixed orifice nozzle would be inevitable. In addition to this problem the major difficulty of the DI was seen as the high mechanical loading, accentuated by the higher smoke-limited fuel/air ratios (A/F) requiring higher boosts to achieve the target rating. While Ricardo’s earlier research work had shown that DI systems could be made to operate up to 4500 rpm naturally aspirated, on balance (see Table 3) the Comet swirl chamber system, developed over many years for the small high-speed commercial engine, was considered to offer greater potential for this particular application. The major problem foreseen was high thermal loading, although the unaided starting and the full multifuel capabilities were also less satisfactory than those of the DI: however, with built-in aids such as could be applied to the multicylinder engine, and with a restriction to CITE fuel, these latter were not considered to be too serious. At the start of the project, then, some consideration was given to the DI as an alternative, and in addition to test running under boosted conditions of an existing 4000 rpm single-cylinder research unit, designs were completed for a DI version of the test engine. Since that date, the increased pressure of noise, smoke, and particularly exhaust emissions legislation, has increasingly favored the divided chamber system, and test work on the DI version is not now likely to take place. ENGINE FRICTION-Comparing the proposed multicylinder high-speed turbocharged engine with a conventional commercial engine of the same cylinder size and number, it was clear that the former would have a significantly higher fmep through the use of higher rotational and mean piston speeds. As already made clear in Table 1,this would pose serious problems in relation to the attainment of both the target output and an acceptable fuel consumption.Fig.2 shows how using a typical commercial engine fmep/speed curve from Ref.1, the estimated fmep of the high-speed multicylinder engine was obtained. In this estimate, some increase in the mechanical friction of the basic engine structure was assumed, since the turbocharged condition would increase cylinder pressures and require larger bearings to give acceptable reliability. In addition, inlet and exhaust pumping losses could add materially to the high-speed fmep, unless acceptable valve sizes could be maintained. By gasoline standards,then,the mechanical efficiency of the unit would be poor,but experience had shown that although attention to detail throughout the design could yield gains,these low levels were implicit in the project specification. SINGLE-CYLINDER TEST ENGINE Based on the considerations outlined, the definitive single-cylinder test engine was designed, the boundary operating conditions for the engine being: bore and stroke,3-1/2 in ф×3-1/2 in (88.9mmф×88.9mm);normal full-load speed range,3000-6000 rpm; and maximum cylinder pressure,2500 lb/in2(17.3Mpa). While the cylinder pressure limit may seem high by conventional standards, past experience has shown the dangers of designing such engines to low limits, and thus inflicting unforeseen limitations on the test program. In fact, originally, with possible work on a DI version in mind, a limit of 3000 lb/in2(20.7Mpa) was set, but as noted, this limit was later reduced for the Comet version. The Comet swirl chamber engine layout is illustrated in Figs.3-5, and the complete engine shown in Fig.6.Of the major components, the following may be said: CRANKCASE-The crankcase and rear-mounted timing case and cover are in gray flake graphite iron to BS 1452:1961 Grade 14,spigoted or doweled and bolted together. The crankcase design was adopted from that of the Ricardo E/6 variable compression ratio gasoline engine, which results in the presence of the front chamber of the crankcase unit, where the E/6 timing drive was situated. Three main bearings are use, all of lead-bronze bushing type, the center bearing being the thrust bearing. The rear bearing acts only as a steady bearing to the otherwise long extension of the crankshaft, the clearance being adjusted so that it cannot take the firing load off the center bearing. CRANKSHAFT-The crankshaft is a one-piece forging in nitriding steel to BS 970:1955 En 40c.The balance weights are integral, and balance only the rotating loads, since primary and secondary balancer shafts are fitted to the engine. All journal and pin surfaces are nitrided, the diameters of the three journals being 3,3,and 2-3/8 in (76.2,76.2,and 60.4 mm),respectively, from front to rear, and of the pin 2-5/8 in (66.6mm). CONNECTING ROD-To obtain the better material properties associated with a forging without the expense of special dies, a search was made of commercial engines, and the connecting rod of the Ford 2700 series diesel engine finally selected as the most appropriate. While satisfactory big-end bearing loadings were achieved at the 3000 lb/in2 maximum cylinder pressures, the little-end design was considered inadequate, and the decision made to use this rod only for the Comet swirl chamber version of the engine, at a pressure limit of 2500 lb/in2. The bearings are as used on the 2700 series engine, that is ,15% reticular tin/aluminium half liners, the little-end bushing being a wrapped lead-bronze item. In addition to careful checking and polishing of the rod, the little end is reduced in width, the better to distribute the firing and inertia loads between the piston pin bosses and the little-end bushing. A higher torque than standard is used on the big-end setscrews, to prevent the cap lifting off due to the inertia forces at tdc exhaust, at 6000 rpm. Computer calculations carried out by the bearing suppliers, The Glacier Metal Co.Ltd., showed that the proposed bearing arrangements were acceptable, although the big end in particular has to accept very arduous conditions at high speeds due to the great inertia of the relatively massive connecting rod (Fig.7).The importance of correct form for both the pin and the bearing under these conditions cannot be overstressed; this apart, the only problem that occurred was rapid cavitation attack in the top (loaded) half liner with the original clearance. The cause of this is evident in Fig.7,and a reduction in clearance to 0.0022 in (56μm) cured this trouble. PISTON AND WRIST PIN-The piston is a one-piece sand casting in 13% silicon aluminum alloy to BS 1490:1970 LM13WP,with the shallow trench and twin recesses of the Comet combustion system formed in one face of the angled (pent-roof) crown surface. Two compression rings are used, the top being a plain barrel-faced ring and the second a taper-faced internally stepped (twisted) ring; the slotted oil-control ring is of the conformable type. Rings are supplied copper-plated on the rubbing faces to assist in bedding in, but are not chrome-plated, since this facing is applied to the liner. The piston was designed deliberately of relatively great height, since it was feared that the very high (by diesel standards) piston speeds together with boosted operation would create difficulties in obtaining acceptable piston, ring, and liner conditions, and it was not thought desirable to accentuate problems more than was necessary. However, relatively little trouble has been experienced with the ring pack. Piston cooling and little-end lubrication is via an oil spray from a fixed jet located in the crankcase. This method was selected to avoid grooving the big-end bearing liner, which would have reduced its capacity. Two piston designs were developed, a tray-cooled arrangement and the soluble core design shown in the figures. To obtain acceptable cooling of the ring belt with the tray-cooled design, the struts transmitting gas loads from the crown to the wrist pin bosses were thinned as far as was thought practicable, but this arrangement was found to allow excessive distortion. The soluble core design has given excellent service to date. The wrist pin is of case-hardened steel,1-3/8 in (34.9mm) in diameter: CYLINDER LINER AND WATER JACKET-The high rates of local heat transfer associated with the use of a swirl chamber combustion system, together with the 2500 lb/in2 maximum cylinder pressure limit, led to design difficulties with the wet-type cylinder liner, since calculations showed that a conventional iron liner of thickness adequate to withstand the gas loads would give excessive surface temperatures for acceptable lubrication at the top ring reversal point. The solution adopted was to use a steel liner, with the bore given the necessary surface finish before being plated with hard chrome to a thickness of 0.0015 in (38μm) by the Chromard process. Toward the top, the liner is thinned to provide the necessary temperature control, while the greater thickness lower down enhances rigidity to combat water-side attack. The liner is flanged at the top and seats on the cylindrical mild steel water jacket itself seating on top of the crankcase; radial location is provided by the liner spigoting in the crankcase, a water seal being obtained in the normal way by rubber O-rings. CYLINDER HEAD ASSEMBLY-The cylinder head, with its associated cambox, is the most complex single assembly of the engine, and presented considerable design problems. The most pressing of these centered on the provision of adequate valves and ports-the difficulties here may be appreciated when it is realized that ports suitable for a conventional engine of 5-1/2 in bore had to be provided on a 3-1/2 in bore-together with adequate cooling for the very high rating of 5.7 ihp/in2 (0.66 indicated kW/cm2) of piston area, this with a swirl chamber comber combustion system. The position of the Comet swirl chamber at the edge of the bore does not render the use of four valves very attractive, and although this and other possible layouts were examined, a 3-valve arrangement was finally adopted. A pent-roof head surface was necessary (Table 2), partly to obtain the necessary valve area but primarily to prevent excessive congestion higher up in the head. It is normally preferable to pair the exhaust valves to reduce their individual size and use a single large valve, but the opposite layout was finally chosen, as shown in Fig.8, since the paired valves had to lie in the center of the head and the intense heating of the head from the long port duct if this latter were the exhaust was considered unacceptable. Asymmetrical chamber/valye layouts were also investigated but rejected as offering no real advantages and being incompatible with multicylinder requirements. The so-called externally inserted form of the Comet chamber was adopted to minimize the space occupied by the hot plug forming the lower portion of the chamber. To cool the resulting four bridges ins the lower deck of the head, between the chamber and the inlet valves, and the inlet and exhaust valves, drillings were provided, giving an accurately controllable metal thickness between the hot gases and the coolant, and a clean surface on the coolant side. The swirl chamber hot plug carrying the throat, and made from Nimonic 80A alloy by precision casting, is a light fit on its sides as well as locating on a copper gasket on its upper flange, since experience showed that at these high ratings some direct cooling was necessary, unlike commercial engines where an air gap is used to improve warm up after starting. The upper part of the chamber carrying the injector is a spheroidal graphite iron casting. A transducer tapping into the cylinder is provided at the front end of the head. The light alloy head seats directly on the steel liner flange, no gasket being employed, and is clamped by eight suds rooted high in the head and passing vertically downward through the water jacket top flange. No difficulties with gas blow have been exper