JP3840871B2

JP3840871B2 - Compression self-ignition gasoline engine

Info

Publication number: JP3840871B2
Application number: JP2000070820A
Authority: JP
Inventors: 博史宮窪; 康治平谷
Original assignee: Nissan Motor Co Ltd
Current assignee: Nissan Motor Co Ltd
Priority date: 2000-03-14
Filing date: 2000-03-14
Publication date: 2006-11-01
Anticipated expiration: 2020-03-14
Also published as: JP2001263067A

Description

【０００１】
【発明の属する技術分野】
本発明は、少なくとも一部の運転領域において、混合気を圧縮して自己着火させる圧縮自己着火式ガソリン機関に関する。
【０００２】
【従来の技術】
圧縮自己着火燃焼は、燃焼室の多点で燃焼が開始されるため燃焼速度が速く、通常の火花点火燃焼に比べて空燃比がリーンな状態でも安定した燃焼を実現することができる。このため、燃料消費率の向上が可能であり、また、空燃比がリーンなため燃焼温度が低下することから、排気ガス中のＮＯｘを大幅に低減することもできる。
【０００３】
また、高回転、高負荷領域では、通常の火花点火燃焼を行わせ、低回転、低中負荷領域では前記火花点火燃焼から圧縮自己着火燃焼に燃焼形態を切り替えることによって、高回転、高負荷時の高出力性能の確保と、低回転、低中負荷時の燃料消費率向上及びＮＯｘの低減という環境性能向上の両立を図ることができる。
【０００４】
ガソリンのような自己着火性の低い燃料を用いて圧縮自己着火燃焼を行なわせる場合、残留ガスの持つ熱エネルギを利用することが有効である。これは例えば特開平１０−２６６８７８号公報に示されているように、排気行程から吸気行程に移行する際に、排気バルブと吸気バルブがともに閉となるマイナスオーバラップ期間（密閉期間）を設けて、残留ガスを積極的に生じさせる所謂内部ＥＧＲを行なわせることで実現される。
【０００５】
【発明が解決しようとする課題】
しかしながら、前記従来の構成にあっては、吸気ポートに噴射された燃料が新気と混合した状態で燃焼室内に導入され、燃焼室内に残留した内部ＥＧＲガスと均一に混合するようになるため、燃焼室壁面近傍の低温層に分布する燃料は圧縮自己着火に至らずに未燃燃料として排出されるので、燃料効率の悪化および排気ＨＣが増加するという問題点があった。
【０００６】
前記燃焼室壁面での未燃燃料は、混合気ガス温度が低い程かつ燃料濃度が低い程増加する傾向を持っているため、低負荷時の圧縮自己着火燃焼ほど顕著に発生する。
【０００７】
本発明はかかる問題点に鑑みたもので、その目的は、燃焼室内における高温のＥＧＲガスおよび燃料分布をコントロールし、排気バルブ近傍の燃焼室内温度を高く、また燃料濃度を濃くすることで、圧縮自己着火燃焼時の未燃燃料の排出を低減するとともに燃費を向上させた圧縮自己着火式ガソリン機関を提供することである。
【０００８】
【課題を解決するための手段】
上記目的を達成するため請求項１記載の発明は、少なくとも一部の運転領域で混合気を圧縮自己着火燃焼させる圧縮自己着火式ガソリン機関において、筒内に直接燃料を噴射する燃料噴射弁を備え、筒内における燃料濃度を排気バルブ側が吸気バルブ側に対して濃くなるように前記燃料噴射弁から燃料噴射することにより、前記圧縮自己着火燃焼時の筒内における燃焼開始位置を排気側としたことを要旨とする。
【００１０】
上記目的を達成するため請求項２記載の発明は、請求項１記載の圧縮自己着火式ガソリン機関において、排気上死点付近で吸気バルブ及び排気バルブが共に閉じたマイナスオーバーラップのバルブタイミングに制御可能な動弁機構を備え、圧縮自己着火運転領域でバルブタイミングを前記マイナスオーバーラップに設定することにより排気の一部を内部ＥＧＲガスとして筒内に残留させることを要旨とする。
【００１１】
上記目的を達成するため請求項３記載の発明は、請求項２記載の圧縮自己着火式ガソリン機関において、前記圧縮自己着火運転領域における筒内のガス温度を排気バルブ側が吸気バルブ側に対して高くなるようにしたことを要旨とする。
【００１２】
上記目的を達成するため請求項４記載の発明は、請求項３記載の圧縮自己着火式ガソリン機関において、燃焼室内のガス流動を制御するガス流動制御手段を吸気系に備え、圧縮自己着火燃焼領域では新気を燃焼室内の吸気バルブ側に分布させ、前記内部ＥＧＲガスを排気側に層状に分布させたことを要旨とする。
【００１３】
上記目的を達成するため請求項５記載の発明は、請求項４記載の圧縮自己着火式ガソリン機関において、前記ガス流動制御手段は、燃焼室内に生成するタンブル流を制御可能なタンブル制御手段であり、圧縮自己着火運転領域では筒内に生成するタンブル流はピストン冠面を吸気側から排気側へ流れる逆タンブル流とし、火花点火燃焼運転領域では前記タンブル流がピストン冠面を排気側から吸気側へ流れる順タンブル流となるように制御することを要旨とする。
【００１４】
上記目的を達成するため請求項６記載の発明は、請求項２ないし請求項４のいずれか１項記載の圧縮自己着火式ガソリン機関において、前記マイナスオーバーラップ期間中に燃料の一部を噴射し、圧縮行程後半に残りの燃料を噴射することを要旨とする。
【００１５】
上記目的を達成するため請求項７記載の発明は、請求項１ないし請求項６のいずれか１項に記載の圧縮自己着火式ガソリン機関において、ピストン冠面に中心軸がクランク軸と並行となる略円筒面状の凹部を有し、該凹部の排気側端部で前記円筒面に接する接平面が、上死点前およそ３０°でのピストン位置において、前記シリンダヘッドと排気バルブの外端より内側で交差するように形成されていることを要旨とする。
【００１６】
上記目的を達成するため請求項８記載の発明は、請求項７に記載の圧縮自己着火式ガソリン機関において、前記燃料噴射弁より圧縮行程中に噴射された燃料が排気バルブの中心付近に偏在するように噴射時期を制御することを要旨とする。
【００１８】
上記目的を達成するため請求項９記載の発明は、シリンダヘッドの燃焼室略中央に点火プラグを有し、かつ筒内に直接燃料を噴射する噴射弁を吸気ポートの下方に配置し、該噴射弁より噴射された燃料により成層混合気を形成したのち火花点火により着火し火炎伝播燃焼する成層燃焼運転領域と、自己着火燃焼運転領域とを有する圧縮自己着火式ガソリン機関において、同一エンジン回転数における自己着火燃焼運転時の圧縮行程後半での噴射時期を前記成層燃焼運転領域での噴射時期よりも早い時期に噴射することを要旨とする。
【００１９】
上記目的を達成するため請求項１０記載の発明は、請求項９に記載の圧縮自己着火式ガソリン機関において、前記自己着火燃焼運転時の圧縮行程後半での噴射時期と圧縮上死点までの間隔が前記成層燃焼運転領域での噴射時期から点火時期までの間隔に対し約２倍となるように設定したことを要旨とする。
【００２０】
上記目的を達成するため請求項１１記載の発明は、圧縮自己着火燃焼と火花点火燃焼とを運転条件により切り換える圧縮自己着火式ガソリン機関において、吸気ポート内部を上下の吸気通路に仕切る仕切り板と、前記上下の吸気通路のいずれか一方を閉じることが可能なタンブル制御弁とを備え、圧縮着火燃焼時には前記タンブル制御弁により上側の吸気通路を閉じて燃焼室内に逆タンブル流を生じさせ、火花点火の成層燃焼時には前記タンブル制御弁により下側の吸気通路を閉じて燃焼室内に順タンブル流を生じさせ、火花点火の均質燃焼時には前記上下の吸気通路を共に開くことを要旨とする。
【００２１】
【発明の効果】
請求項１記載の発明によれば、筒内における燃料濃度を排気バルブ側が吸気バルブ側に対して濃くなるようにしたため、圧縮自己着火燃焼が筒内の排気側から開始されるようになり、排気バルブ近傍の燃焼室壁面における燃料は高温な既燃ガスに曝されかつ燃焼の進行とともに圧縮を受けるため高温となり易く燃焼が進み、吸気バルブ側に対し未燃燃料の発生が抑えられる。排気行程では燃焼度が高く未燃ＨＣが殆どない排気バルブ側のガスが先に排出され、吸気バルブ側に存在し筒内未燃ＨＣの大半を占めるガスが自己ＥＧＲとして筒内に残留しやすいため、排気ガス中に存在する未燃ＨＣが低減し熱効率向上と排気の浄化が図れる。
【００２２】
さらにシリンダヘッドの燃焼室壁面における温度分布は排気バルブ付近が高温であること、および排気バルブ付近では濃い混合気が存在することにより排気バルブ近傍の未燃燃料が低減する。
【００２３】
請求項２記載の発明によれば、排気上死点付近で吸気バルブ及び排気バルブが共に閉じたマイナスオーバーラップのバルブタイミングに制御可能な動弁機構を備え、圧縮自己着火運転領域でバルブタイミングを前記マイナスオーバーラップに設定することにより、吸気側に存在する未燃燃料を含むガスは前記マイナスオーバーラップにより排気バルブが排気行程途中で閉となるためさらに排出され難くなり、残留ガスとともに次サイクルに持ち越されることとなる。よって請求項１に記載の効果をより高めることができる。
【００２４】
請求項３記載の発明によれば、筒内のガス温度分布を排気バルブ側が吸気バルブ側に対して高くなるようにしているため、排気側からの燃焼開始を実現できるとともに、排気バルブ付近での混合気温度が上昇するため排気バルブ近傍の未燃燃料が低減し、請求項１または２に記載の効果をより高めることができる。
【００２５】
請求項４記載の発明によれば、燃焼室内のガス流動を制御するガス流動制御手段を吸気系に備え、圧縮自己着火燃焼領域では新気を燃焼室内の吸気バルブ側に分布させ、前記内部ＥＧＲガスを排気側に層状に分布させたことにより、請求項３記載の排気側が吸気側より高い筒内ガス温度分布を実現できる。
【００２６】
請求項５記載の発明によれば、前記ガス流動制御手段は、燃焼室内に生成するタンブル流を制御可能なタンブル制御手段であり、圧縮自己着火運転領域では筒内に生成するタンブル流はピストン冠面を吸気側から排気側へ流れる逆タンブル流とし、火花点火燃焼運転領域では前記タンブル流がピストン冠面を排気側から吸気側へ流れる順タンブル流となるように制御するようにしたので、圧縮自己着火燃焼時の筒内温度分布と、火花点火燃焼時の筒内ガス流動とを両立させ、火花点火運転領域の性能を犠牲にすることなく、請求項４記載の効果を得ることが可能となる。
【００２７】
請求項６記載の発明によれば、前記マイナスオーバーラップ期間中に燃料の一部を燃焼室内に供給するようにしているため、前記マイナスオーバーラップ期間中の上死点近傍において前記内部ＥＧＲ中に残存する酸素と未燃燃料および新たに投入した前記燃料が部分的に酸化され、反応性の高い燃料に改質される。この改質された燃料はより低温域での燃焼が可能となるため、請求項２から４の効果に加えて、燃焼室壁面での未燃燃料の発生を抑制することが可能となる。
【００２８】
請求項７記載の発明によれば、ピストン冠面に中心軸がクランク軸と並行となる略円筒面状の凹部を有し、該凹部の排気側端部で前記円筒面に接する接平面が、上死点前およそ３０°のピストン位置において、前記シリンダヘッドと排気バルブの外端より内側で交差するように形成されている。このため燃料噴射弁より供給された燃料は前記ピストン冠面の凹部に沿ってシリンダヘッド面の中でも高温な排気バルブ近傍へと輸送されるため、請求項１から６記載の効果を確実に得ることが可能となる。
【００２９】
請求項８記載の発明によれば、燃料噴射弁から噴射された燃料が排気バルブの中心を指向するタイミングに噴射するため、請求項７の効果を確実に得ることが可能となる。
【００３１】
請求項９および１０記載の発明によれば、圧縮自己着火運転領域での圧縮行程における燃料噴射時期を適切に設定することができるため排気ガス中に含まれる未燃燃料を低減することが可能となる。
【００３２】
請求項１１記載の発明によれば、吸気ポート内部を上下の吸気通路に仕切る仕切り板と、前記上下の吸気通路のいずれか一方を閉じることが可能なタンブル制御弁とを備え、圧縮着火燃焼時には前記タンブル制御弁により上側の吸気通路を閉じて燃焼室内に逆タンブル流を生じさせ、火花点火の成層燃焼時には前記タンブル制御弁により下側の吸気通路を閉じて燃焼室内に順タンブル流を生じさせ、火花点火の均質燃焼時には前記上下の吸気通路を共に開くようにしたので、圧縮自己着火燃焼時に排気側の温度を高めた筒内温度分布と、火花点火成層燃焼時の筒内ガス流動と、火花点火均質燃焼時の性能とをそれぞれ満足させた圧縮自己着火式ガソリン機関を提供することができる。
【００３３】
【発明の実施の形態】
以下、図面に基づいて本発明の実施の形態について説明する。
図１は、本発明に係る圧縮自己着火式ガソリン機関の第１の実施形態を示すシステム構成図である。図１において、シリンダブロック１と、ピストン２と、シリンダヘッド３とにより燃焼室４が形成されている。
【００３４】
シリンダヘッド３には、吸気ポート５とこれら吸気ポート５を開閉する吸気バルブ６、および吸気ポート５と対向的に配置された排気ポート７とこれら排気ポート７を開閉する排気バルブ８を備えている。
【００３５】
吸気バルブ６と排気バルブ８は、それぞれ吸気カム９と排気カム１０を介して図外のバルブ駆動系により開閉される。このバルブ駆動系はエンジンコントロールユニット２２からの指示により開閉時期可変手段１１，１２を介して吸気バルブ６、排気バルブ８の開閉時期を制御可能な構成としてある。即ち、機関の低、中負荷領域では実質的な圧縮比の変更、内部ＥＧＲガス量などを制御し、圧縮自己着火運転が可能な高温、高圧状態を実現できる構成としている。
【００３６】
吸気ポート５の上流には吸気管１３が接続されており、吸気管１３には下流側端面近くにタンブル制御バルブ１４を付設してあり、上流側には空気量調整用スロットルバルブ１５と図示しない空気量測定用のエアフロメーター、エアクリーナ等を設けてある。
【００３７】
さらに吸気ポート５の内部には、吸気バルブ６の近傍から吸気管１３との接合面に渡って、管を上下に分割する仕切り板１８ａが設けられている。同様に吸気管１３には吸気ポート５との接合面からタンブル制御バルブ１４の回転中心軸に渡る仕切り板１８ｂが前記仕切り板１８ａと連続的に設けられている。
【００３８】
タンブル制御バルブ１４とスロットルバルブ１５は、それぞれエンジンコントロールユニット２２により開閉手段１６，１７を介してバルブ開閉制御可能としてある。
【００３９】
一方、シリンダヘッド３には吸気ポート５の下に臨んで、燃料ポンプ２３から供給されるガソリン燃料を直接燃焼室４内に噴射する燃料噴射弁１９を設けてある。
【００４０】
またシリンダヘッド３には燃焼室４内の略中心位置に点火プラグ２０が設けられており、点火プラグ２０は主に高回転、高負荷時に通常の火花点火燃焼を行なう場合に使用する。
【００４１】
エンジンコントロールユニット２２には、機関運転条件を示す信号として、機関回転数信号、クランク角度信号、負荷信号、空気量信号、吸気温度信号、排気温度信号、燃圧信号、油水温信号などが入力され、これら各種の信号に基づいて演算処理を実施し、前記吸気バルブ６、排気バルブ８のバルブタイミング、タンブル制御バルブ１４、スロットルバルブ１５の各バルブ開度制御、燃料噴射弁１９の噴射量と噴射時期、および点火プラグ２０の点火時期を適切に制御している。
【００４２】
図２は、ピストン２の冠面形状を示す平面図及び側面図である。図２に示すように、ピストン２の冠面略中央部には機関のクランク軸と略平行な中心軸を有する略円筒面状の凹部２４を設けてある。そして、ピストン２の凹部２４の排気側の端部で前記円筒面に接する接平面２５が、上死点前およそ３０°でのピストン位置において、前記シリンダヘッド３と排気バルブ８の外端２６より内側で交差するように形成されている。
【００４３】
このため、凹部２４により吸気行程で形成されたタンブル流が圧縮行程まで維持され、さらに圧縮行程後半で燃料噴射弁１９より噴射された燃料が凹部２４に沿って吸気側から排気側に輸送され、圧縮上死点付近では排気バルブ８付近に到達するようになっている。
【００４４】
図９の（ａ）、（ｂ）は吸気バルブ６、排気バルブ８のバルブタイミングの可変制御の一例を示しており、火花点火運転時には通常のバルブタイミングである（ａ）の状態とし、排気上死点付近において排気バルブ８と吸気バルブ６がともに開となる所定量のバルブオーバーラップ時期が発生するように設定される。
【００４５】
圧縮自己着火運転時にはバルブタイミングを（ｂ）の状態とし、すなわち排気バルブ８の閉時期が進角して排気行程途中で閉弁するとともに、吸気バルブ６の開時期が遅角して吸気行程途中で開弁するように制御されマイナスオーバーラップ状態になるように設定されている。
【００４６】
このように排気上死点付近で吸排気弁が共に閉じたマイナスオーバーラップ期間（密閉期間）を成すバルブタイミングとすることで、排気バルブ閉時期の燃焼室容積に相当する既燃ガスを燃焼室４内に滞留させて次サイクルへ内部ＥＧＲガスとして持ち越すことが可能となる。この内部ＥＧＲガスのもつ熱エネルギを有効に利用して後述するようにリーン空燃比での圧縮自己着火燃焼が圧縮上死点付近で実現される。
【００４７】
次に、本実施形態の動作について説明する。
図３は、各運転モードにおけるタンブル制御バルブ１４の開閉状態と、燃焼室４内に形成されるタンブル流の状態を説明する図である。
【００４８】
圧縮自己着火運転時には、図３（ａ）に示すように、仕切り板１８で仕切られた吸気ポート５の上段５ａの通路が閉となり、同下段５ｂの通路が開となる状態にタンブル制御バルブ１４を制御する。この結果、新気は主に吸気バルブ６の外側を通って燃焼室４内に流入し、燃焼室４内には弱い逆タンブル流２７が発生する。そのため燃焼室４内では冷えた新気が吸気側およびピストン２の上面に主に分布し、他方、前述のマイナスオーバーラップにより閉じ込められていた内部ＥＧＲガスは相対的に排気バルブ７の付近に分布し、筒内のガス温度は排気バルブ７付近がより高温な状態となる。
【００４９】
一方、低回転での火花点火運転時では、図３（ｂ）に示すように、吸気ポート５の下段５ｂの通路が閉となり、同上段５ａの通路が開となる状態にタンブル制御バルブ１４を制御する。これにより燃焼室４内には強い順タンブル流２８が発生し、燃料噴射弁１９からの燃料噴射タイミングが吸気行程前半の場合には均一な混合気生成を助けるとともに、噴射タイミングが圧縮行程後半となるような場合には噴霧貫徹力と順タンブル流による流動成分があいまって、点火プラグ２０付近に成層混合気を形成することができ、火花点火での成層燃焼を実現する。
【００５０】
高回転、高負荷での火花点火運転時では、図３（ｃ）に示すように、タンブル制御バルブ１４は仕切り板１８と略平行となる状態に制御し、吸気ポート５の上段５ａ及び下段５ｂが共に開いた状態にする。このため新気の流入抵抗を増加させることがなく、火花点火運転時の全負荷時の性能が維持される。
【００５１】
次に、図４を用いて本実施形態の燃料分布状態を説明する。図４（ａ）は、圧縮行程前半における燃焼室内の新気とＥＧＲガスの分布を示し、図４（ｂ）は、圧縮行程後半の燃料噴射の様子を示し、図４（ｃ）は、圧縮行程後半の燃料拡散の様子をそれぞれ示すものである。
【００５２】
図４（ｂ）に示すように、圧縮自己着火運転時の燃料は、圧縮行程後半に前記ピストン冠面に形成した凹部２４を指向して噴射される。該燃料は気化、拡散しながら自らの貫徹力および燃焼室４内に生成している逆タンブル流により、前記凹部２４の円筒面に沿って排気バルブ側に輸送され、混合気の燃料濃度は排気バルブ近傍が濃く、吸気側に向かう程薄くなる状態に形成される。
【００５３】
ピストン冠面の前記凹部２４の排気側端部で前記円筒面に接する接平面２５が、上死点前およそ３０°でのピストン位置において、シリンダヘッド３と排気バルブ８の外端２６より内側で交差するように形成されているため、燃料が低温の燃焼室壁面に囲まれた図４中のＡ部に進入することを抑制している。
【００５４】
さらに燃料の噴射時期としては、圧縮上死点付近で混合気の燃料濃度の中心が排気バルブ８の中心となるように設定している。これにより燃料はシリンダヘッド３の燃焼室４内に臨む面の内、もっとも高温となる排気バルブ８近傍に分布することとなる。
【００５５】
この結果、図６に示すように燃焼室壁面近傍のガス温度低温域で発生する未燃ガスは、排気バルブ８近傍において、内部ＥＧＲの選択的分布、混合気濃度のリッチ化、および高温の燃焼室壁温の効果が協働することにより大幅に減少する。
【００５６】
次に、図５を参照して自己着火燃焼時の着火から膨張行程にかけての状態を説明する。
【００５７】
燃焼室４内の混合気は排気バルブ近傍がリッチかつ高温であるため、最初の着火は、図５（ａ）に示すように排気バルブ８付近で発生し、ここで燃焼したガスが膨張することで、その周囲の未燃状態の混合気は排気バルブ８に対向する燃焼室面に向かって圧縮されて行く。この圧縮により未燃状態の混合気温度が上昇し、燃料は図５（ｂ）に示すように順次着火、燃焼して行く。
【００５８】
燃焼終了時には、図５（ｃ）に示すように、ピストン冠面、および吸気バルブ付近にのみ着火に至らなかった未燃燃料の層が存在することとなる。
【００５９】
以上の結果、図８に示すように、排気ポート７で計測される排気バルブ８開時に発生する未燃燃料（ＨＣ）濃度の第１のピークは、排気バルブ８付近の燃焼室壁面で発生する未燃燃料を低減したことで低下し、排気行程終了前に発生する第２のピーク、すなわち吸気側燃焼室壁面、ピストン冠面等で発生した未燃燃料は排出される前に、排気バルブ８が閉となるため、内部ＥＧＲガス中に閉じ込められ、次サイクルにおいて燃焼することとなる。よって、燃焼効率の向上と排気ガスの浄化を同時に得ることを可能としている。
【００６０】
上記効果に加え、前記マイナスオーバーラップ期間中に燃料の一部を噴射すると、排気上死点付近での圧縮により、内部ＥＧＲガス中に存在する酸素を用いた燃料改質が可能である。自己着火運転領域の低負荷域では、着火性の確保および相対的にリーン混合気での自己着火燃焼を行なう必要があるが、前記燃料改質により、内部ＥＧＲガス中の燃料を改質することで、リーンな混合気においても排気バルブ８付近で発生する未燃燃料を低減することが可能となる。
【００６１】
また、自己着火燃焼領域における圧縮行程後半における噴射時期と排気ポート７で計測されるＨＣ排出率との関係の一例を図１１に示す。自己着火燃焼領域におけるＨＣ排出率の低減を図るためには、最適な噴射時期が存在することがわかる。これは早すぎる噴射時期では燃料の拡散が進み過ぎるため混合気が希薄となり、反対に遅過ぎる噴射時期では排気バルブ８付近に充分濃い混合気を形成できず、図６に示す未燃燃料低減効果が充分には得られなくなるためである。
【００６２】
前記最適な噴射時期は、同一のエンジン回転数において適合された火花点火、成層燃焼運転領域の噴射時期よりも早いことがわかる。これは前記成層燃焼運転領域の噴射時期は成層混合気がシリンダヘッド３の略中央に配置された点火プラグ２０に到達するのに適した時期であるのに対し、圧縮自己着火運転領域の噴射時期は前記点火プラグ２０を通り越し排気バルブ８付近に到達しなければならず、長い時間を要するためである。
【００６３】
さらに、圧縮自己着火燃焼時の最適噴射時期は、噴射から圧縮上死点までに要する期間が、同一エンジン回転数における前記成層燃焼時の噴射から点火までの期間の約２倍となる時期となる。これは、圧縮自己着火燃焼時の着火は圧縮上死点付近で発生すること、および燃料噴射弁１９から排気バルブ８までの距離は該燃料噴射弁１９から点火プラグ２０までの距離の２倍弱であるが、燃料噴霧の持つ貫徹力が減衰することとあいまって生じるためである。
【００６４】
次に図１０に基づいて本発明の第２実施の形態について説明する。
第２実施形態では、第１実施形態に対し、第２点火プラグ２１を排気ポート７の下方で燃焼室４内に臨む位置に追加して設置したことを特徴としている。第２点火プラグ２１はエンジンコントロールユニット２２により点火時期が制御されており、主に圧縮自己着火燃焼時に放電を行い圧縮自己着火燃焼開始を補助するために用いている。
【００６５】
第２実施形態では第１実施形態の効果に加えて、以下の効果を得ることができる。低負荷域で混合気濃度が相対的に薄い場合では、第２点火プラグ２１による放電を圧縮上死点前４０°付近で実施し、排気バルブ８近傍にラジカル（活性化学種）を生成させることで、リーン混合気であっても確実な着火を得ることができる。
【００６６】
また、高負荷域では急速な燃焼を回避するために圧縮上死点以降に着火をさせる必要がある。このような状況では圧縮だけでは上死点前に着火が発生しないように前記マイナスオーバーラップ量を調整し、かつ第２点火プラグ２１で上死点以降での放電を実施することで要求する時期に着火を発生させることが可能となる。
【図面の簡単な説明】
【図１】本発明に係る圧縮自己着火式ガソリン機関の第１実施形態の構成図である。
【図２】実施形態におけるピストン冠面形状を示した図である。
【図３】実施形態の各運転モードにおけるタンブル制御バルブの状態と燃焼室内に形成されるタンブル流の状態を示した図である。
【図４】実施形態における圧縮行程の混合気状態を示した図である。
【図５】実施形態における膨張行程の混合気状態を示した図である。
【図６】燃焼室内ガスの温度分布と未燃燃料発生の状況を示した図である。
【図７】実施形態におけるタンブル制御バルブの効果を示した図である。
【図８】本発明による未燃燃料低減の効果を示した図である。
【図９】本発明における吸、排気バルブタイミングの設定の一例を示す図である。
【図１０】本発明の第２実施形態の要部を示す図である。
【図１１】本発明における噴射時期と排気ＨＣ排出率の関係を示した図である。
【符号の説明】
１シリンダブロック
２ピストン
３シリンダヘッド
４燃焼室
５吸気ポート
６吸気バルブ
７排気ポート
８排気バルブ
９吸気カム
１０排気カム
１１吸気バルブ開閉時期可変手段
１２排気バルブ開閉時期可変手段
１３吸気管
１４タンブル制御バルブ
１５スロットルバルブ
１６タンブル制御バルブ開閉手段
１７スロットルバルブ開閉手段
１８仕切り板
１９燃料噴射弁
２０点火プラグ
２１第２点火プラグ
２２エンジンコントロールユニット
２３燃料ポンプ
２４ピストン冠面凹部
２５凹部接平面
２６排気バルブ外端[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a compression self-ignition gasoline engine that compresses and self-ignites an air-fuel mixture in at least a part of the operation region.
[0002]
[Prior art]
In the compression self-ignition combustion, combustion is started at many points in the combustion chamber, so that the combustion speed is high, and stable combustion can be realized even when the air-fuel ratio is lean as compared with normal spark ignition combustion. Therefore, the fuel consumption rate can be improved, and the NOx in the exhaust gas can be greatly reduced because the combustion temperature is lowered because the air-fuel ratio is lean.
[0003]
In addition, normal spark ignition combustion is performed in the high rotation and high load regions, and in the low rotation and low and medium load regions, the combustion mode is switched from the spark ignition combustion to the compression self-ignition combustion. It is possible to achieve both high output performance and improved environmental performance such as low fuel consumption rate improvement at low rotation and low and medium loads, and NOx reduction.
[0004]
When compression self-ignition combustion is performed using a low self-ignition fuel such as gasoline, it is effective to use the thermal energy of the residual gas. For example, as disclosed in Japanese Patent Laid-Open No. 10-266878, a negative overlap period (sealing period) is provided in which both the exhaust valve and the intake valve are closed when the exhaust stroke is shifted to the intake stroke. This is realized by performing so-called internal EGR that actively generates residual gas.
[0005]
[Problems to be solved by the invention]
However, in the conventional configuration, the fuel injected into the intake port is introduced into the combustion chamber in a state of being mixed with fresh air, and is uniformly mixed with the internal EGR gas remaining in the combustion chamber. Since the fuel distributed in the low temperature layer in the vicinity of the wall surface of the combustion chamber does not reach compression self-ignition and is discharged as unburned fuel, there is a problem that fuel efficiency is deteriorated and exhaust HC increases.
[0006]
The unburned fuel on the wall of the combustion chamber has a tendency to increase as the gas mixture temperature is lower and the fuel concentration is lower, so that it is more prominent as the compression self-ignition combustion at low load.
[0007]
The present invention has been made in view of such problems, and its purpose is to control the distribution of high-temperature EGR gas and fuel in the combustion chamber, increase the temperature in the combustion chamber near the exhaust valve, and increase the fuel concentration. It is an object of the present invention to provide a compression self-ignition gasoline engine that reduces the emission of unburned fuel during self-ignition combustion and improves fuel efficiency.
[0008]
[Means for Solving the Problems]
  In order to achieve the above-mentioned object, the invention according to claim 1 is a compression self-ignition gasoline engine in which an air-fuel mixture is subjected to compression self-ignition combustion in at least a part of an operation region.A fuel injection valve that directly injects fuel into the cylinder, and by injecting fuel from the fuel injection valve so that the exhaust valve side is thicker than the intake valve side in the cylinder,The gist is that the combustion start position in the cylinder at the time of the compression self-ignition combustion is the exhaust side.
[0010]
  Claims to achieve the above object2The described invention is claimed.1The compression self-ignition gasoline engine is equipped with a valve mechanism that can be controlled to a negative overlap valve timing in which both the intake valve and the exhaust valve are closed near the exhaust top dead center, and the valve timing is adjusted in the compression self-ignition operation region. The gist is that a part of the exhaust gas remains in the cylinder as internal EGR gas by setting the minus overlap.
[0011]
  Claims to achieve the above object3The described invention is claimed.2The gist of the compression self-ignition gasoline engine is that the gas temperature in the cylinder in the compression self-ignition operation region is higher on the exhaust valve side than on the intake valve side.
[0012]
  Claims to achieve the above object4The described invention is claimed.3In the compression self-ignition gasoline engine described above, gas flow control means for controlling gas flow in the combustion chamber is provided in the intake system, and in the compression self-ignition combustion region, fresh air is distributed to the intake valve side in the combustion chamber, and the internal EGR The gist is that the gas is distributed in layers on the exhaust side.
[0013]
  Claims to achieve the above object5The described invention is claimed.4In the compression self-ignition gasoline engine described above, the gas flow control means is a tumble control means capable of controlling a tumble flow generated in the combustion chamber. In the compression self-ignition operation region, the tumble flow generated in the cylinder is a piston crown. The gist is to control the surface to be a reverse tumble flow that flows from the intake side to the exhaust side, and in the spark ignition combustion operation region, the tumble flow is a forward tumble flow that flows from the exhaust side to the intake side.
[0014]
  Claims to achieve the above object6The described invention is claimed.2 to 4In the compression self-ignition gasoline engine according to any one of the above, the gist is that a part of the fuel is injected during the minus overlap period and the remaining fuel is injected in the latter half of the compression stroke.
[0015]
  Claims to achieve the above object7The invention described in claim 1 to claim 16The compression self-ignition gasoline engine according to any one of the above, wherein the piston crown surface has a substantially cylindrical concave portion whose central axis is parallel to the crankshaft, and the cylindrical surface at the exhaust side end of the concave portion The tangent plane in contact with the cylinder is formed so as to intersect the cylinder head and the outer end of the exhaust valve at the piston position at about 30 ° before top dead center.
[0016]
  Claims to achieve the above object8The described invention is claimed.7In the compression self-ignition gasoline engine described in 1), the gist is to control the injection timing so that the fuel injected from the fuel injection valve during the compression stroke is unevenly distributed near the center of the exhaust valve.
[0018]
  Claims to achieve the above object9In the described invention, an ignition valve having an ignition plug in the center of the combustion chamber of the cylinder head and injecting fuel directly into the cylinder is disposed below the intake port, and stratified mixing is performed by the fuel injected from the injection valve. In the compression self-ignition type gasoline engine having a stratified combustion operation region that is ignited by spark ignition and then flame propagation combustion is formed, and a self-ignition combustion operation region, the latter half of the compression stroke during the self-ignition combustion operation at the same engine speed The gist of the present invention is to inject at a timing earlier than the injection timing in the stratified combustion operation region.
[0019]
  Claims to achieve the above object10The described invention is claimed.9In the compression self-ignition gasoline engine described in 1., the interval between the injection timing in the latter half of the compression stroke and the compression top dead center during the self-ignition combustion operation is the interval from the injection timing to the ignition timing in the stratified combustion operation region. The gist is that it is set to be about double.
[0020]
  Claims to achieve the above object11In the compression self-ignition gasoline engine that switches between compression self-ignition combustion and spark ignition combustion according to operating conditions, the invention described in any one of the partition plate that divides the interior of the intake port into upper and lower intake passages and one of the upper and lower intake passages A tumble control valve capable of closing the engine, and when the compression ignition combustion is performed, the upper intake passage is closed by the tumble control valve to generate a reverse tumble flow in the combustion chamber, and during stratified combustion of the spark ignition, the tumble control valve The gist is to close the lower intake passage to generate a forward tumble flow in the combustion chamber, and to open the upper and lower intake passages together during homogeneous ignition with spark ignition.
[0021]
【The invention's effect】
  According to invention of Claim 1,Because the fuel concentration in the cylinder is made higher on the exhaust valve side than on the intake valve side,Compression self-ignition combustion starts from the exhaust side in the cylinderLikeSince the fuel on the wall of the combustion chamber near the exhaust valve is exposed to high-temperature burned gas and is compressed as the combustion progresses, the fuel tends to reach a high temperature and combustion proceeds, and the generation of unburned fuel on the intake valve side is suppressed. In the exhaust stroke, the exhaust valve side gas having a high burnup and almost no unburned HC is discharged first, and the gas existing on the intake valve side and occupying most of the unburned HC in the cylinder is likely to remain in the cylinder as self EGR. Therefore, unburned HC present in the exhaust gas is reduced, so that the thermal efficiency can be improved and the exhaust gas can be purified.
[0022]
  Furthermore, the temperature distribution on the combustion chamber wall surface of the cylinder head reduces unburned fuel near the exhaust valve due to the high temperature in the vicinity of the exhaust valve and the presence of a rich mixture near the exhaust valve..
[0023]
  Claim2According to the described invention, the valve timing mechanism that can be controlled to a negative overlap valve timing in which both the intake valve and the exhaust valve are closed near the exhaust top dead center is provided, and the valve timing is reduced in the compression self-ignition operation region. By setting the lap, the gas containing unburned fuel on the intake side is more difficult to be discharged because the exhaust valve is closed during the exhaust stroke due to the minus overlap, and it is carried over to the next cycle together with the residual gas. It becomes. So claims1The described effects can be further enhanced.
[0024]
  Claim3According to the described invention, because the gas temperature distribution in the cylinder is higher on the exhaust valve side than on the intake valve side, combustion can be started from the exhaust side, and the mixture temperature in the vicinity of the exhaust valve can be realized. So that unburned fuel near the exhaust valve is reduced, andOr 2The described effects can be further enhanced.
[0025]
  Claim4According to the described invention, the gas flow control means for controlling the gas flow in the combustion chamber is provided in the intake system, fresh air is distributed to the intake valve side in the combustion chamber in the compression self-ignition combustion region, and the internal EGR gas is exhausted. Claimed by being distributed in layers on the side3The in-cylinder gas temperature distribution which is higher on the exhaust side than the intake side can be realized.
[0026]
  Claim5According to the described invention, the gas flow control means is a tumble control means capable of controlling the tumble flow generated in the combustion chamber, and in the compression self-ignition operation region, the tumble flow generated in the cylinder sucks the piston crown surface. Compressed self-ignition combustion because the reverse tumble flow that flows from the exhaust side to the exhaust side is controlled and the tumble flow is controlled to be the forward tumble flow that flows from the exhaust side to the intake side in the spark ignition combustion operation region. In-cylinder temperature distribution at the same time and in-cylinder gas flow at the time of spark ignition combustion, and without sacrificing the performance of the spark ignition operation region,4The described effects can be obtained.
[0027]
  Claim6According to the described invention, since a part of the fuel is supplied into the combustion chamber during the minus overlap period, oxygen remaining in the internal EGR near the top dead center during the minus overlap period. The unburned fuel and the newly introduced fuel are partially oxidized and reformed into a highly reactive fuel. Since this reformed fuel can be burned at lower temperatures,2 to 4In addition to the above effect, it is possible to suppress the generation of unburned fuel on the combustion chamber wall surface.
[0028]
  Claim7According to the described invention, the piston crown surface has a substantially cylindrical concave portion whose central axis is parallel to the crankshaft, and the tangential plane in contact with the cylindrical surface at the exhaust side end of the concave portion is top dead center. The cylinder head and the exhaust valve are formed so as to intersect inside the outer end of the exhaust valve at a piston position of approximately 30 ° in front. For this reason, since the fuel supplied from the fuel injection valve is transported along the concave portion of the piston crown surface to the vicinity of the high-temperature exhaust valve in the cylinder head surface.6The described effects can be obtained with certainty.
[0029]
  Claim8According to the described invention, the fuel injected from the fuel injection valve is injected at a timing directed to the center of the exhaust valve.7It is possible to reliably obtain the effect.
[0031]
  Claim9 and 10According to the described invention, since the fuel injection timing in the compression stroke in the compression self-ignition operation region can be set appropriately, it is possible to reduce unburned fuel contained in the exhaust gas.
[0032]
  Claim11According to the described invention, it is provided with the partition plate that partitions the inside of the intake port into the upper and lower intake passages, and the tumble control valve capable of closing either one of the upper and lower intake passages, and the tumble control is performed at the time of compression ignition combustion. The upper intake passage is closed by a valve to create a reverse tumble flow in the combustion chamber. During stratified combustion by spark ignition, the lower intake passage is closed by the tumble control valve to produce a forward tumble flow in the combustion chamber, and spark ignition is performed. Since the upper and lower intake passages are opened at the time of homogeneous combustion, the in-cylinder temperature distribution in which the temperature on the exhaust side is increased during compression self-ignition combustion, the in-cylinder gas flow at the time of spark ignition stratified combustion, and the homogeneous ignition of spark ignition It is possible to provide a compression self-ignition gasoline engine that satisfies the performance during combustion.
[0033]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, embodiments of the present invention will be described with reference to the drawings.
FIG. 1 is a system configuration diagram showing a first embodiment of a compression self-ignition gasoline engine according to the present invention. In FIG. 1, a combustion chamber 4 is formed by a cylinder block 1, a piston 2, and a cylinder head 3.
[0034]
The cylinder head 3 includes an intake port 5, an intake valve 6 that opens and closes the intake port 5, an exhaust port 7 that is disposed opposite to the intake port 5, and an exhaust valve 8 that opens and closes the exhaust port 7. .
[0035]
The intake valve 6 and the exhaust valve 8 are opened and closed by a valve drive system (not shown) via an intake cam 9 and an exhaust cam 10, respectively. This valve drive system is configured to be able to control the opening / closing timing of the intake valve 6 and the exhaust valve 8 via the opening / closing timing varying means 11, 12 according to an instruction from the engine control unit 22. In other words, in the low and medium load regions of the engine, a substantial change in compression ratio, internal EGR gas amount, and the like are controlled to realize a high temperature and high pressure state capable of compression self-ignition operation.
[0036]
An intake pipe 13 is connected upstream of the intake port 5, and a tumble control valve 14 is provided near the downstream end face of the intake pipe 13, and an air amount adjusting throttle valve 15 is not shown on the upstream side. An air flow meter, air cleaner, etc. are provided for measuring the amount of air.
[0037]
Further, a partition plate 18 a that divides the pipe vertically is provided in the intake port 5 from the vicinity of the intake valve 6 to the joint surface with the intake pipe 13. Similarly, a partition plate 18b extending continuously from the joint surface with the intake port 5 to the rotation center axis of the tumble control valve 14 is provided on the intake pipe 13 continuously with the partition plate 18a.
[0038]
The tumble control valve 14 and the throttle valve 15 can be controlled to be opened / closed by the engine control unit 22 via the opening / closing means 16 and 17, respectively.
[0039]
On the other hand, the cylinder head 3 is provided with a fuel injection valve 19 that faces under the intake port 5 and injects gasoline fuel supplied from the fuel pump 23 directly into the combustion chamber 4.
[0040]
The cylinder head 3 is provided with a spark plug 20 at a substantially central position in the combustion chamber 4. The spark plug 20 is mainly used when normal spark ignition combustion is performed at high speed and high load.
[0041]
An engine speed signal, a crank angle signal, a load signal, an air amount signal, an intake air temperature signal, an exhaust gas temperature signal, a fuel pressure signal, an oil / water temperature signal, etc. are input to the engine control unit 22 as signals indicating engine operating conditions. Arithmetic processing is performed based on these various signals to control the valve timing of the intake valve 6 and the exhaust valve 8, the valve opening control of the tumble control valve 14 and the throttle valve 15, the injection amount and the injection timing of the fuel injection valve 19. And the ignition timing of the spark plug 20 are appropriately controlled.
[0042]
FIG. 2 is a plan view and a side view showing the crown shape of the piston 2. As shown in FIG. 2, a substantially cylindrical concave portion 24 having a central axis substantially parallel to the crankshaft of the engine is provided at a substantially central portion of the crown surface of the piston 2. The tangential plane 25 that contacts the cylindrical surface at the exhaust-side end of the recess 24 of the piston 2 is more than the cylinder head 3 and the outer end 26 of the exhaust valve 8 at the piston position at approximately 30 ° before top dead center. It is formed so as to intersect inside.
[0043]
For this reason, the tumble flow formed in the intake stroke by the recess 24 is maintained until the compression stroke, and the fuel injected from the fuel injection valve 19 in the latter half of the compression stroke is transported from the intake side to the exhaust side along the recess 24. Near the compression top dead center, the exhaust valve 8 is reached.
[0044]
FIGS. 9A and 9B show an example of variable control of the valve timings of the intake valve 6 and the exhaust valve 8, and during the spark ignition operation, the normal valve timing is set to the state (a), and the exhaust timing is increased. In the vicinity of the dead point, a predetermined amount of valve overlap timing is set at which both the exhaust valve 8 and the intake valve 6 are opened.
[0045]
During the compression self-ignition operation, the valve timing is set to the state (b), that is, the closing timing of the exhaust valve 8 is advanced to close in the middle of the exhaust stroke, and the opening timing of the intake valve 6 is retarded to be in the middle of the intake stroke. The valve is controlled so as to open at a negative overlap state.
[0046]
Thus, by setting the valve timing to the minus overlap period (sealing period) in which the intake and exhaust valves are both closed near the exhaust top dead center, the burned gas corresponding to the combustion chamber volume at the exhaust valve closing timing is discharged to the combustion chamber. It is possible to retain the gas in the gas and carry it over to the next cycle as internal EGR gas. As will be described later, the compression self-ignition combustion at the lean air-fuel ratio is realized in the vicinity of the compression top dead center by effectively using the thermal energy of the internal EGR gas.
[0047]
Next, the operation of this embodiment will be described.
FIG. 3 is a diagram illustrating the open / closed state of the tumble control valve 14 and the state of the tumble flow formed in the combustion chamber 4 in each operation mode.
[0048]
During the compression self-ignition operation, as shown in FIG. 3A, the tumble control valve 14 is brought into a state in which the passage of the upper stage 5a of the intake port 5 partitioned by the partition plate 18 is closed and the passage of the lower stage 5b is opened. To control. As a result, fresh air mainly flows into the combustion chamber 4 through the outside of the intake valve 6, and a weak reverse tumble flow 27 is generated in the combustion chamber 4. Therefore, in the combustion chamber 4, the cool fresh air is mainly distributed on the intake side and the upper surface of the piston 2, while the internal EGR gas confined by the above-described minus overlap is relatively distributed near the exhaust valve 7. However, the temperature of the gas in the cylinder is higher in the vicinity of the exhaust valve 7.
[0049]
On the other hand, during the spark ignition operation at low rotation, as shown in FIG. 3B, the tumble control valve 14 is set so that the passage of the lower stage 5b of the intake port 5 is closed and the passage of the upper stage 5a is opened. Control. As a result, a strong forward tumble flow 28 is generated in the combustion chamber 4. When the fuel injection timing from the fuel injection valve 19 is in the first half of the intake stroke, uniform air-fuel mixture generation is assisted and the injection timing is in the second half of the compression stroke. In such a case, a stratified mixture can be formed in the vicinity of the spark plug 20 by combining the spray penetration force and the flow component due to the forward tumble flow, thereby realizing stratified combustion by spark ignition.
[0050]
At the time of spark ignition operation at high speed and high load, as shown in FIG. 3C, the tumble control valve 14 is controlled so as to be substantially parallel to the partition plate 18, and the upper stage 5a and the lower stage 5b of the intake port 5 are controlled. Are both open. For this reason, the inflow resistance of fresh air is not increased, and the performance at the full load during the spark ignition operation is maintained.
[0051]
Next, the fuel distribution state of this embodiment will be described with reference to FIG. 4A shows the distribution of fresh air and EGR gas in the combustion chamber in the first half of the compression stroke, FIG. 4B shows the state of fuel injection in the second half of the compression stroke, and FIG. It shows the state of fuel diffusion in the second half of the process.
[0052]
As shown in FIG. 4B, the fuel during the compression self-ignition operation is injected toward the recess 24 formed in the piston crown surface in the latter half of the compression stroke. The fuel is transported to the exhaust valve side along the cylindrical surface of the recess 24 by its penetration force and the reverse tumble flow generated in the combustion chamber 4 while vaporizing and diffusing, and the fuel concentration of the mixture is reduced to the exhaust gas. The vicinity of the valve is dark and is formed so as to become thinner toward the intake side.
[0053]
A tangential plane 25 in contact with the cylindrical surface at the exhaust side end of the concave portion 24 on the piston crown surface is located on the inner side of the cylinder head 3 and the outer end 26 of the exhaust valve 8 at the piston position at approximately 30 ° before top dead center. Since it is formed so as to intersect, the fuel is prevented from entering the portion A in FIG. 4 surrounded by the low temperature combustion chamber wall surface.
[0054]
Further, the fuel injection timing is set so that the center of the fuel concentration of the air-fuel mixture becomes the center of the exhaust valve 8 near the compression top dead center. As a result, the fuel is distributed in the vicinity of the exhaust valve 8 having the highest temperature in the surface of the cylinder head 3 facing the combustion chamber 4.
[0055]
As a result, as shown in FIG. 6, the unburned gas generated in the low temperature region near the combustion chamber wall surface has a selective distribution of internal EGR, enrichment of the air-fuel mixture concentration, and high temperature combustion in the vicinity of the exhaust valve 8. The effect of room wall temperature is greatly reduced by cooperation.
[0056]
Next, referring to FIG. 5, the state from the ignition to the expansion stroke at the time of self-ignition combustion will be described.
[0057]
Since the air-fuel mixture in the combustion chamber 4 is rich and high in the vicinity of the exhaust valve, the first ignition occurs in the vicinity of the exhaust valve 8 as shown in FIG. 5A, and the burned gas expands here. The surrounding unburned air-fuel mixture is compressed toward the combustion chamber surface facing the exhaust valve 8. This compression raises the temperature of the unburned mixture, and the fuel is ignited and burned sequentially as shown in FIG.
[0058]
At the end of the combustion, as shown in FIG. 5C, an unburned fuel layer that has not been ignited exists only on the piston crown and in the vicinity of the intake valve.
[0059]
As a result of the above, as shown in FIG. 8, the first peak of unburned fuel (HC) concentration generated when the exhaust valve 8 is measured, which is measured at the exhaust port 7, occurs on the combustion chamber wall near the exhaust valve 8. The exhaust valve 8 is reduced before the second peak generated before the end of the exhaust stroke, that is, the unburned fuel generated on the intake side combustion chamber wall surface, the piston crown surface, etc. is discharged. Is closed, so that it is trapped in the internal EGR gas and burned in the next cycle. Therefore, improvement in combustion efficiency and purification of exhaust gas can be obtained at the same time.
[0060]
In addition to the above effects, when a part of fuel is injected during the minus overlap period, fuel reforming using oxygen present in the internal EGR gas is possible by compression near the exhaust top dead center. In the low-load region of the self-ignition operation region, it is necessary to ensure ignitability and to perform self-ignition combustion with a lean air-fuel mixture, but reforming the fuel in the internal EGR gas by the fuel reforming Thus, it is possible to reduce the unburned fuel generated in the vicinity of the exhaust valve 8 even in a lean air-fuel mixture.
[0061]
An example of the relationship between the injection timing in the latter half of the compression stroke in the self-ignition combustion region and the HC emission rate measured at the exhaust port 7 is shown in FIG. It can be seen that there is an optimal injection timing in order to reduce the HC emission rate in the self-ignition combustion region. This is because the fuel diffusion is too advanced at the injection timing that is too early and the mixture becomes lean. On the other hand, at the injection timing that is too late, a sufficiently rich mixture cannot be formed in the vicinity of the exhaust valve 8, and the unburned fuel reduction effect shown in FIG. This is because it cannot be obtained sufficiently.
[0062]
It can be seen that the optimum injection timing is earlier than the injection timing in the spark ignition and stratified combustion operation region adapted at the same engine speed. This is because the injection timing in the stratified charge combustion operation region is a time suitable for the stratified mixture to reach the spark plug 20 disposed substantially at the center of the cylinder head 3, whereas the injection timing in the compression self-ignition operation region. This is because it takes a long time to pass through the spark plug 20 and reach the vicinity of the exhaust valve 8.
[0063]
Furthermore, the optimum injection timing at the time of compression self-ignition combustion is a time when the period required from injection to compression top dead center is approximately twice the period from injection to ignition at the stratified combustion at the same engine speed. . This is because the ignition at the time of compression self-ignition combustion occurs near the compression top dead center, and the distance from the fuel injection valve 19 to the exhaust valve 8 is slightly less than twice the distance from the fuel injection valve 19 to the spark plug 20. However, this is because the penetration force of the fuel spray is combined with the attenuation.
[0064]
Next, a second embodiment of the present invention will be described with reference to FIG.
The second embodiment is characterized in that the second spark plug 21 is additionally provided at a position facing the combustion chamber 4 below the exhaust port 7 with respect to the first embodiment. The ignition timing of the second spark plug 21 is controlled by the engine control unit 22. The second spark plug 21 is mainly used for discharging during compression self-ignition combustion and assisting the start of compression self-ignition combustion.
[0065]
In the second embodiment, the following effects can be obtained in addition to the effects of the first embodiment. When the air-fuel mixture concentration is relatively low in the low load region, discharge by the second spark plug 21 is performed near 40 ° before compression top dead center, and radicals (active chemical species) are generated in the vicinity of the exhaust valve 8. Thus, reliable ignition can be obtained even with a lean air-fuel mixture.
[0066]
In addition, it is necessary to ignite after compression top dead center in order to avoid rapid combustion in a high load range. In such a situation, the timing required by adjusting the minus overlap amount so that ignition does not occur before top dead center only by compression and performing discharge after top dead center with the second spark plug 21. It becomes possible to cause ignition to occur.
[Brief description of the drawings]
FIG. 1 is a configuration diagram of a first embodiment of a compression self-ignition gasoline engine according to the present invention.
FIG. 2 is a view showing a piston crown shape in the embodiment.
FIG. 3 is a diagram showing a state of a tumble control valve and a state of a tumble flow formed in a combustion chamber in each operation mode of the embodiment.
FIG. 4 is a diagram showing an air-fuel mixture state in a compression stroke in the embodiment.
FIG. 5 is a diagram showing an air-fuel mixture state in an expansion stroke in the embodiment.
FIG. 6 is a diagram showing the temperature distribution of combustion chamber gas and the state of unburned fuel generation.
FIG. 7 is a diagram showing an effect of a tumble control valve in the embodiment.
FIG. 8 is a diagram showing the effect of reducing unburned fuel according to the present invention.
FIG. 9 is a diagram showing an example of setting of intake and exhaust valve timings in the present invention.
FIG. 10 is a diagram showing a main part of a second embodiment of the present invention.
FIG. 11 is a graph showing the relationship between the injection timing and the exhaust HC emission rate in the present invention.
[Explanation of symbols]
1 Cylinder block
2 piston
3 Cylinder head
4 Combustion chamber
5 Intake port
6 Intake valve
7 Exhaust port
8 Exhaust valve
9 Intake cam
10 Exhaust cam
11 Intake valve opening / closing timing variable means
12 Exhaust valve opening / closing timing variable means
13 Intake pipe
14 Tumble control valve
15 Throttle valve
16 Tumble control valve opening / closing means
17 Throttle valve opening / closing means
18 Partition plate
19 Fuel injection valve
20 Spark plug
21 Second spark plug
22 Engine control unit
23 Fuel pump
24 Piston crown recess
25 Concave surface of recess
26 Exhaust valve outer end

Claims

In a compression self-ignition gasoline engine in which the air-fuel mixture is subjected to compression self-ignition combustion in at least a part of the operation region,
A fuel injection valve for directly injecting fuel into the cylinder, and by injecting fuel from the fuel injection valve so that the exhaust valve side is thicker than the intake valve side in the cylinder; A compression self-ignition gasoline engine characterized in that the combustion start position in the cylinder is on the exhaust side.

Equipped with a valve mechanism that can be controlled to a negative overlap valve timing in which both the intake valve and exhaust valve are closed near the exhaust top dead center, and exhausting by setting the valve timing to the negative overlap in the compression self-ignition operation region 2. The compression self-ignition gasoline engine according to claim 1, wherein a part of the engine is left as an internal EGR gas in the cylinder.

The compression self-ignition gasoline engine according to claim 2, wherein the gas temperature in the cylinder in the compression self-ignition operation region is set higher on the exhaust valve side than on the intake valve side.

A gas flow control means for controlling the gas flow in the combustion chamber is provided in the intake system, and in the compression self-ignition combustion region, fresh air is distributed on the intake valve side in the combustion chamber, and the internal EGR gas is distributed in layers on the exhaust side. The compression self-ignition gasoline engine according to claim 3.

The gas flow control means is a tumble control means capable of controlling the tumble flow generated in the combustion chamber. In the compression self-ignition operation region, the tumble flow generated in the cylinder is reversely flowing from the intake side to the exhaust side on the piston crown surface. 5. The compression self-ignition gasoline engine according to claim 4, wherein the tumble flow is controlled so that the tumble flow becomes a forward tumble flow flowing from the exhaust side to the intake side in the spark ignition combustion operation region. .

The compressed self-ignition gasoline engine according to any one of claims 2 to 4, wherein a part of the fuel is injected during the minus overlap period, and the remaining fuel is injected in the latter half of the compression stroke. .

A piston having a substantially cylindrical surface-shaped recess whose central axis is parallel to the crankshaft on the piston crown surface, and a tangential plane that is in contact with the cylindrical surface at the exhaust side end of the recess is approximately 30 ° before top dead center The compression self-ignition gasoline engine according to any one of claims 1 to 6, wherein the compression self-ignition gasoline engine is formed so as to intersect with the cylinder head at an inner side than an outer end of the exhaust valve.

8. The compression self-ignition gasoline engine according to claim 7, wherein the injection timing is controlled so that fuel injected from the fuel injection valve during the compression stroke is unevenly distributed near the center of the exhaust valve.

After having an ignition plug in the center of the combustion chamber of the cylinder head and injecting fuel directly into the cylinder below the intake port, the fuel injected from the injection valve forms a stratified mixture In a compression self-ignition gasoline engine having a stratified combustion operation region ignited by spark ignition and flame propagation combustion, and a self-ignition combustion operation region, the injection timing in the latter half of the compression stroke at the time of self-ignition combustion operation at the same engine speed A compression self-ignition gasoline engine, which is injected at a time earlier than an injection time in the stratified combustion operation region.

The interval between the injection timing and the compression top dead center in the latter half of the compression stroke during the self-ignition combustion operation is set to be approximately twice the interval from the injection timing to the ignition timing in the stratified combustion operation region. The compression self-ignition gasoline engine according to claim 9.

In a compression self-ignition gasoline engine that switches between compression self-ignition combustion and spark ignition combustion according to operating conditions, it is possible to close either the partition plate that partitions the intake port interior into upper and lower intake passages or the upper and lower intake passages A tumble control valve for closing the upper intake passage by the tumble control valve at the time of compression ignition combustion and generating a reverse tumble flow in the combustion chamber, and the lower intake passage by the tumble control valve at the time of stratified combustion by spark ignition A compression self-ignition gasoline engine characterized by closing the valve to generate a forward tumble flow in the combustion chamber and opening both the upper and lower intake passages at the time of homogeneous combustion with spark ignition.