JPS588453A - Deceleration sensing-fluid pressure controller of vehicle brake system - Google Patents
Deceleration sensing-fluid pressure controller of vehicle brake systemInfo
- Publication number
- JPS588453A JPS588453A JP10431481A JP10431481A JPS588453A JP S588453 A JPS588453 A JP S588453A JP 10431481 A JP10431481 A JP 10431481A JP 10431481 A JP10431481 A JP 10431481A JP S588453 A JPS588453 A JP S588453A
- Authority
- JP
- Japan
- Prior art keywords
- spring
- hydraulic pressure
- valve
- force
- fluid pressure
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Pending
Links
Classifications
-
- B—PERFORMING OPERATIONS; TRANSPORTING
- B60—VEHICLES IN GENERAL
- B60T—VEHICLE BRAKE CONTROL SYSTEMS OR PARTS THEREOF; BRAKE CONTROL SYSTEMS OR PARTS THEREOF, IN GENERAL; ARRANGEMENT OF BRAKING ELEMENTS ON VEHICLES IN GENERAL; PORTABLE DEVICES FOR PREVENTING UNWANTED MOVEMENT OF VEHICLES; VEHICLE MODIFICATIONS TO FACILITATE COOLING OF BRAKES
- B60T8/00—Arrangements for adjusting wheel-braking force to meet varying vehicular or ground-surface conditions, e.g. limiting or varying distribution of braking force
- B60T8/26—Arrangements for adjusting wheel-braking force to meet varying vehicular or ground-surface conditions, e.g. limiting or varying distribution of braking force characterised by producing differential braking between front and rear wheels
- B60T8/28—Arrangements for adjusting wheel-braking force to meet varying vehicular or ground-surface conditions, e.g. limiting or varying distribution of braking force characterised by producing differential braking between front and rear wheels responsive to deceleration
- B60T8/282—Arrangements for adjusting wheel-braking force to meet varying vehicular or ground-surface conditions, e.g. limiting or varying distribution of braking force characterised by producing differential braking between front and rear wheels responsive to deceleration using ball and ramp
Landscapes
- Engineering & Computer Science (AREA)
- Transportation (AREA)
- Mechanical Engineering (AREA)
- Hydraulic Control Valves For Brake Systems (AREA)
Abstract
Description
【発明の詳細な説明】
本発明は車両ブレーキ系の減速度感知式液圧制御装置の
改良、特に後輪ブレーキ液圧の減圧制御の折点値を可変
制御させる機構に関するものである。DETAILED DESCRIPTION OF THE INVENTION The present invention relates to an improvement in a deceleration sensing type hydraulic pressure control device for a vehicle brake system, and particularly to a mechanism for variably controlling the corner value of pressure reduction control of rear wheel brake hydraulic pressure.
従来より、車両ブレーキ力はその制動時における前後車
輪の路面への押付力の差異から、前輪に比べて後輪側を
所定の割合で低減させる必要のあることはよく知られて
おり、またこの押付力も車両の荷重積載状態の変化に応
じて変化するものであることも知られている・
そこでこのような変化に対応し、出来るだけ理想的な前
後輪デレーキカの配分を行なりて車輪々、りの発生を防
止する工夫が種々なされており、例えば入・出力液圧室
に臨む液圧受圧面積の異なるピストンと、仁のピストン
にバネ力を付勢するスプリングとの設定関係により、後
輪プレー中波圧【折点減圧制御するデｏ、３１１−シ、
ニンダ型の液圧制御弁と、車両制動時の減速度が一定値
に達したときに封止液室への連通を遮断して、この封止
された液圧ｇＩＫ応じた状態に前記スプリングを圧縮さ
せ、前記ピストンへの液圧作用力とバネ方のバランスに
よる折点減圧制御の開始点を可変させるようにした慣性
弁吟のＧパルプとを併せ備えたものが提供されている＠
仁のような形式の液圧制御装置は、車両の荷重積載の増
大に伴なって一定減速＆を得るに要するブν−キカが、
比例的関係をもって増大することから、前後ブレーキ力
の理想配分に近似した制御特性を得るようＫなされたも
のである・ところで、この工うな折点減圧制御の開始点
（以下液圧折点と称する）の可変制御は、車両の荷重積
載量の増大に略比例的な関係ｔなすといっても、空車状
ＷＡＫ近いときと積載量が限界（定積と称する）Ｋ近い
ときとでは、その液圧折点の上昇の割・８は一定ではな
く、望車に近い軽積載ｗ＃には上昇割合を小さく、反対
に定積に近い重積戦時には大きくすることが、理想的な
ブレーキ方配分比に近似するものであるとされる・
本発明は、このような点に鑑み、車両の荷重積載の状態
に略比例的な関係で、一定の車両減速度を得るＫｌ！す
るブレーキ液圧値が上昇し、したがって封止液圧も同じ
く略比例的に上昇する構成の液圧制御装置において、プ
ロ、／−シ、ニング型の液圧制御弁に作用するバネ方の
可変制御は、前記封止液圧を受けて単に比例的に増大さ
せるのではなく、ある一定の段階（軽積載がら重積載の
中間の状態における段階）で増大率を大きく変えるよう
にしたものであり、具体的に本発明の要旨は、ブレーキ
液圧をマスクシリンダがら後輪ブレーキ装置Ｋ伝見る径
路に介設され、液圧作用と付勢バネ方の／４ランスにて
後輪ブレーキ液圧を折点減圧制御するプロボーシ、二ン
グ聾液圧制御弁と、車両制動時の減速度が一定ｗｉを越
えたときのブレーキ液圧を封止液ＷｉＫ封止するＧパル
プとを備え、この封止液圧値に対し略比例的にバネカ會
増大するスデリンダａｔｓｔ介して前記液圧制御弁への
付勢バネ力を増大させ液圧折点値管増大させる形式の車
両ブレーキ系の減速贋感知式液圧制御装置において、前
記スｆ １７ンダ４１１！構は、液圧制御弁にΔネカを
付勢する主たるスプリングと、この主たるスプリングに
抗したバネ方を作用しうる２本の調整用スプリングとｔ
Ｖし、前記封止ｔｉａの液圧を受けて移動するように設
けられた調整ピストンの移動量が一定値を越えるときの
前後で、前記２本の調整用スプリングの／臂ネ力作用の
保合が切換わりて増大するよう構成したこと１−特徴と
する車両ブレーキ系の減速度感知式液圧制御量にある。It has been well known that vehicle braking force needs to be reduced by a predetermined ratio on the rear wheels compared to the front wheels due to the difference in the pressing force of the front and rear wheels against the road surface during braking. It is also known that the pushing force changes depending on the changes in the loading condition of the vehicle. Therefore, in response to such changes, the front and rear wheels are distributed as ideally as possible. Various measures have been taken to prevent this from occurring. For example, by setting the relationship between the pistons facing the input and output hydraulic pressure chambers with different hydraulic pressure receiving areas and the spring that applies spring force to the inner piston, the rear wheel Wave pressure during play [break point pressure reduction control deo, 311-shi,
When the deceleration during braking of the vehicle reaches a certain value, communication with the sealing liquid chamber is cut off using a Ninda-type hydraulic pressure control valve, and the spring is placed in a state corresponding to the sealed liquid pressure gIK. There is a product that combines the G-pulp of the inertia valve, which is compressed and the starting point of corner pressure reduction control is varied by the balance between the hydraulic force acting on the piston and the spring side. In this type of hydraulic control device, the pressure ν-kika required to obtain a constant deceleration & is increased as the load carrying capacity of the vehicle increases.
Since the hydraulic pressure increases in a proportional relationship, it was designed to obtain control characteristics that approximate the ideal distribution of front and rear brake forces.By the way, this technique is designed to obtain control characteristics that approximate the ideal distribution of front and rear brake forces. Although the variable control of ) is approximately proportional to the increase in the vehicle's load carrying capacity, the difference in fluid flow between when the car is close to the empty WAK and when the payload is close to the limit (referred to as constant volume) K. The rate of rise at the breaking point is not constant, and the ideal brake distribution is to reduce the rate of increase when the car is lightly loaded near the desired vehicle, and to increase it when the load is close to a constant load. In view of these points, the present invention is designed to obtain a constant vehicle deceleration in a relationship substantially proportional to the load state of the vehicle. In a hydraulic pressure control device configured such that the brake fluid pressure value increases, and therefore the sealing fluid pressure also increases approximately proportionally, the spring type that acts on the hydraulic pressure control valve of the pressure control valve is variable. The control does not simply increase proportionally in response to the sealing fluid pressure, but greatly changes the rate of increase at a certain stage (stage between light loading and heavy loading). Specifically, the gist of the present invention is that the mask cylinder is interposed in the path through which the brake fluid pressure is transmitted to the rear wheel brake device K, and the rear wheel brake fluid pressure is transmitted through the hydraulic action and the quarter lance of the biasing spring. Equipped with a hydraulic pressure control valve that performs pressure reduction control at the turning point, and a G pulp that seals the brake fluid pressure when the deceleration during vehicle braking exceeds a certain value WiK, this sealing A deceleration counterfeit detection type fluid for a vehicle brake system that increases a biasing spring force to the hydraulic pressure control valve through a Sdelinda atst which increases the spring force approximately proportionally to the hydraulic pressure value, thereby increasing the hydraulic pressure corner value. In the pressure control device, the above f 17 nd 411! The structure consists of a main spring that biases the hydraulic pressure control valve with a Δ force, two adjustment springs that can act in a manner that resists this main spring, and t.
V, and before and after the amount of movement of the adjusting piston, which is provided to move in response to the hydraulic pressure of the sealing tia, exceeds a certain value, the elastic force of the two adjusting springs is maintained. The present invention is characterized by a deceleration-sensing hydraulic pressure control amount of a vehicle brake system.
以下本発明の実施態様を図面に基づいて説明する。Embodiments of the present invention will be described below based on the drawings.
尚図面に示す実施例は本発明を二重配管用の滅遮匿感知
式淑圧制御装置に適用した場合のものであるが、−重用
のものでもよいことは轟然である・１１ＫＴｈイ？１
、２は１４にブーｚｒ４．３．４．５．６はゾロ４−シ
、ニンｒ作動渥の液圧制御弁を収容すゐ段付のＶＩＪ：
／〆であり、その最大径シリン／６（図の右側）Ｏ開口
はスジリンダ収１１１１７に臨んでいる。The embodiment shown in the drawings is a case where the present invention is applied to a non-blocking sensing type suction pressure control device for double piping, but it is of course possible to apply it to a device for heavy-duty use. 1
Stepped VIJ:
/〆, and its maximum diameter cylinder /6 (right side in the figure) O opening faces the suji cylinder housing 11117.
前記液圧制御弁は、既知の二重配管用の液圧制御機構を
なしてシＣ１５Ｆｉ筒状の７．イルセイフピストン、９
はバランスシリンダ、１０はバックアップ、１１＃ｉ中
シリンダ部材、１２は中シリン〆、１３はバランスシリ
ンダ９に滑合されたバランスピストンであり、その一端
は１系入方液寵ｂ１を挿通して出方液＊ｂ、に臨み、他
端はム系田ヵｆ［ｌ！ ａ ｍ Ｋ臨んでいる・１４は
フェイルセイフピストン８０内筒部に組付ケラれたバル
ブシートであり、これがバランスピストン１３の頭部に
形成された弁体部と協働し０系弁部をなしている。The hydraulic pressure control valve constitutes a known hydraulic pressure control mechanism for double piping and is a C15Fi cylindrical 7. Il-safe piston, 9
is a balance cylinder, 10 is a backup, 11 #i middle cylinder member, 12 is a middle cylinder end, 13 is a balance piston slidably fitted to the balance cylinder 9, one end of which is inserted through the 1 system inlet fluid port b1. The exit fluid *b, faces the other end, and the other end is the mu system field kaf [l! a m K ・14 is a valve seat assembled into the inner cylinder part of the fail-safe piston 80, and this cooperates with the valve body part formed in the head of the balance piston 13 to operate the 0 system valve part. I am doing it.
１５はホールドスプリング、１ｌｌｔｊ係止リングであ
る。15 is a hold spring, and 1lltj is a locking ring.
１７はＡ系の制御ピストンで６９、−熾はム系出力液ａ
ｌａｓ 円で７エイル七イフクリツ７’１８１１パ２
ンスビス）７１３と一定長の離反限界を持つよう連結さ
れ、他端は中シリンダ１２ｆ挿通してヌデリンダ収容部
７に臨んでいる。１Ｇはシリンダ５内に組付けられたノ
５゛プシ一トであり、制御ピストン１７０頭部に形成さ
れ九弁体部と協働してム系弁部ｔなしている・ ２０は
ホールトスｌリンダ、２１．２２はスジリンダ塵、２３
はピストンカップ、２４は係止りンダである・また、ス
プリング収容１１７には、制御ピストン１７會−原図の
左側端）方向に抑圧する九めの組合せスプリング機構が
収容されてお〕、このスプリング機構は、制御ピストン
１７の他蝋図の右側端）と対向するように形成・配置さ
れた調整シリンダ３１、シよびこれに滑合Ｏ調整ピスト
ン３２によって、鋏調整ピストン３２の臨む封止液１１
Ｃの封止液圧の増大に応じて圧縮されバネ力管増大する
ように設けられている４０であるが、本実施例の特徴は
このスプリング機構の構成およびその動作に基づ〈作用
・効果Ｋ１１ｌｉ徽があ）、この点については詳ｌｌｌ
Ｋ１＆述する。17 is the A system control piston 69, - is the mu system output liquid a
las yen de 7 ael 7 ifcrits 7'1811 par 2
The other end of the inner cylinder 12f passes through the middle cylinder 12f and faces the cylinder housing portion 7. 1G is a nozzle assembled in the cylinder 5, which is formed on the head of the control piston 170 and cooperates with the nine valve body parts to form a valve body T. 20 is a hole cylinder. , 21.22 is sujilinda dust, 23
24 is a piston cup, and 24 is a locking member. Also, the spring housing 117 houses a ninth combined spring mechanism that suppresses the control piston 17 in the direction (the left end in the original drawing), and this spring mechanism The adjustment cylinder 31 is formed and arranged to face the control piston 17 (the right end in the figure), and the sealing fluid 11 facing the scissor adjustment piston 32 is slidably fitted thereto.
The spring mechanism 40 is compressed and the spring force increases as the sealing fluid pressure increases. K11li Hui), more details on this point
K1&describe.
このような構成をなす液圧制御弁の作動について簡単に
述べると、通常図示する静止位置にある各ピストンによ
って、ム・Ｂ系入力液部ａｌ、ｂＩＫマスタシリンダよ
り液圧が伝えられると、この液圧はＡ−Ｂ系弁１１１１
を通じて出力液部１３．ｂ−次いで夫々の後輪ブレーキ
装置に伝えられる−そして制御ピストン１７の入出力筐
璽＆ｔｓ＆ｍＫ臨む液圧受圧面積の大小関係と、前記ス
デリンダ機構Ｏ付勢バネカとＯａ係で、ム系は一定のル
−キ液圧値から出カ濱圧Ｐ、を入力液圧Ｐ、に対し折点
減圧制御を始め、これに伴うてＢ系の折点減圧制御が受
動的に生ずることになる。To briefly describe the operation of the hydraulic pressure control valve with such a configuration, when hydraulic pressure is transmitted from the M/B system input liquid parts al and bIK master cylinder by each piston normally in the rest position shown in the figure, this Hydraulic pressure is A-B system valve 1111
Through the output liquid part 13. b - Next, it is transmitted to each rear wheel brake device - and due to the magnitude relationship of the hydraulic pressure receiving area facing the input/output housing &ts&mK of the control piston 17, and the Sderinda mechanism O biasing spring and Oa, the MU system is maintained at a constant level. Corner point pressure reduction control is started with respect to the input hydraulic pressure P by converting the output pressure P from the rookie hydraulic pressure value, and along with this, corner point pressure reduction control of the B system is passively generated.
そしてこのときの折点値は調整ピストン３２によるスプ
リング機構のバネカ増大Ｏ８Ｉｍ！によ多可変増大され
るが、この調整ピストン３２０移動は後記する慣性弁に
おける封止液量Ｃ内の封止液圧Ｐ、にて定まることにな
る１、尚ψｌｌ夫夫の入力１ｉｉｉ町、ｂｓは夫々同列
台系の賃スタシリンダに連通され、出力＊ｉｉす、ｂｓ
は夫々各基の後Ｖレーキ装置に連通されておｐ１シたが
うて前輪側に比べて後輪側におけるブレーキ力の低減が
得す１ことになる・次ｒＫ慣性弁にりいて説明すると、
３３は慣性弁収容部であシ、パルプがディ１ 會２０締
結固定Ｋ ｊ Ｏて筒状のシート部＃３４と、弁座３８
管含む弁座構成体３７を位置決めする。シート部材３４
円のゴール収容部は弁座の開口を介して封止液寵ＣＫ連
通されている・３５はシート部材３４の内面１ｓＫ形成
された慣性−−ル３（１０ガイド面でＴｏ夛、図の矢印
に示す車両進行に対して仰角−ｔなしておｐ１制動時の
車両減速度が一定値管越えたと＠に慣性が一ル３６が弁
座３８Ｋｌｌ座して封止液量Ｃを封止するよう構成され
て−る・そしてζＯ封止液寵Ｃがｉｉ＊シリｙｓ”ｓｘ
内Ｏ脚整ピストン３２の端部に臨まれる構成をなしてい
る。The break point value at this time is the increase in spring force of the spring mechanism due to the adjustment piston 32 O8Im! However, the movement of the adjusting piston 320 is determined by the sealing fluid pressure P in the sealing fluid amount C in the inertia valve, which will be described later. bs are respectively communicated with the cylinders in the same series, and the outputs *ii and bs
are communicated with the rear V rake device of each unit, so that it is advantageous to reduce the braking force on the rear wheels compared to the front wheels.Explaining the following rK inertia valve: ,
33 is an inertial valve accommodating part, and the pulp is connected to the cylindrical seat part #34 and the valve seat 38.
Position the valve seat structure 37 including the tube. Sheet member 34
The circular goal accommodating portion is connected to the sealing liquid supply port CK through the opening of the valve seat. ・35 is an inertia rule 3 formed on the inner surface of the seat member 34 (10 on the guide surface, indicated by the arrow in the figure) When the vehicle deceleration during braking exceeds a certain value at an elevation angle of -t with respect to the vehicle progress shown in , the inertia causes the valve seat 38 to sit and seal the sealing fluid amount C. It is composed of - and ζO sealing liquid C is ii* series ys"sx
It is configured to face the end of the inner O-leg adjustment piston 32.
また、慣性弁収容部３３とシート部材３４の外周との間
に形成された周状の間隙が、ム系出液室島雪と４−ト（
図示せず）を介した後輪ブレーキ装置との間の通液路３
９を構成し、更に、この通液路３９とシート部材３４筒
内の慣性が−ル収容部との間の連通をなす通液開口４０
．４１が、骸慣性−−ル３６の移動前後において、対を
なすよう離隔形成されている。In addition, the circumferential gap formed between the inertial valve accommodating portion 33 and the outer periphery of the seat member 34 is connected to the four-tooth (
(not shown) to the rear wheel brake device.
9, and further includes a liquid passage opening 40 that communicates between this liquid passage 39 and the inertia tube housing portion in the cylinder of the sheet member 34.
．． 41 are formed in pairs before and after the movement of the skeleton inertia lever 36 so as to be separated from each other.
このような構成により慣性弁における慣性＆ −ル３６
の弁座３８への烏合着座によりて封止液室ＣＯ筐圧値が
定まり、これにより後輪ツレーキ装置への折点減圧制御
され九液圧も定まることになるという基本的な作動関係
は既存のものと同様である。１＊、本例では慣性が一ル
３６を収容しているシート部材３４内への液伝達が対を
なす通液開口４０．４１を介して行なわれる九めに、液
流は過ｉｌｌ開口４０次いで４１の順に慣性が一ル収容
部内に伝えられるものとなる。このような伝達の一後紘
実際上極めて微小時間の差で紘あゐが、これがツレーキ
操作の急緩と相関して慣性が一ル３６の移動に棗好な効
果をも九らすことＫなり、例えば比較的ブレーキ液圧の
立ち上９が遅い通常のブレーキ時にシいては、一対の通
液開口４０゜４１から流入される液流による慣性が−ル
３６への影響は相互に打消し合って慣性力以外の要素に
よる誤動作の虞れを防止するが、他方急ブレーキ時等の
ブレーキ液圧の立ち上りが速い場合には、系内でマスタ
シリンダ儒に位置する過液開口４゜からの流入液流の影
響が大きくなり、ためにこの液流の影響を受けて慣性−
−ル３６の弁座３８に烏合する移動傾向を増幅し、通常
の慣性移動のみでは弁座３８の開口の閉塞が遅れて異常
に高い液にを封止することがある弊害を防止することが
できるという利点もある。With such a configuration, the inertia &-rule 36 in the inertia valve is reduced.
The basic operating relationship is that the sealing fluid chamber CO casing pressure value is determined by the seating on the valve seat 38, which controls the corner point depressurization of the rear wheel brake device and also determines the hydraulic pressure. It is similar to that of . 1*, in this example, the liquid flow into the sheet member 34, which houses the inertial tube 36, takes place via a pair of liquid passage openings 40, 41; Next, inertia is transmitted into the container in the order of 41. After such transmission, there is actually an extremely small time difference, but this correlates with the sudden and slowing of the rake operation and has a positive effect on the movement of the inertia. For example, during normal braking, where the rise in brake fluid pressure is relatively slow, the influence of inertia on the loop 36 due to the fluid flow flowing in from the pair of fluid passage openings 40 and 41 cancels each other out. This prevents the possibility of malfunction due to factors other than inertial force, but on the other hand, if the brake fluid pressure rises rapidly during sudden braking, the leakage from the excess fluid opening 4° located at the master cylinder pressure in the system The influence of the inflowing liquid flow becomes large, and the inertia is affected by this liquid flow.
- Amplifying the movement tendency of the valve seat 38 of the valve seat 36 to prevent the problem that normal inertial movement alone may delay the closing of the opening of the valve seat 38 and seal it to an abnormally high level of liquid. There is also the advantage of being able to do so.
なお、本実施例では、封止原意Ｃの空気抜１１゜九めに
、弁座構成体３７、とツリーダールトの構成により、空
気抜き時において自動的に慣性−一ル３６の弁座ｓ８へ
の場合が係止されるよ−に１にされている。即ち、弁座
構成体３７は、シート部材３４を介してパルｆがディ１
．２の締結により固定される段付筒状のプラダ４２と、
このｆ２ダ４２の段付内筒部所定位置に後記する軸体の
押付は力で固定的に係止されるシールホルダ４３と、こ
のシールホル〆４３とグラ、グ４２によりて固定される
前記弁座３８とからなり、前記シールホルダ〆４１には
、慣性が−ル３６を収容しているシート部材３４の内部
と封止液室Ｃの間を連通する軸心ＳＯＷ通路４４が形成
され、また弁座３８はこの液通路４４の慣性弁側開口よ
り若干突出した位置で、慣性Ｉ−ルと蟲合しうるように
設置されている。In addition, in this embodiment, at the 11° ninth position for air venting in the sealing principle C, due to the configuration of the valve seat structure 37 and the tree dart, the valve seat s8 of the inertial valve 36 is automatically moved when air is vented. It is set to 1 so that the case is locked. That is, the valve seat structure 37 allows the pulse f to pass through the seat member 34.
．． A stepped cylindrical Prada 42 fixed by fastening 2;
The shaft body (described later) is pressed to a predetermined position of the stepped inner cylinder part of the f2 cylinder 42 by a seal holder 43 that is fixedly locked by force, and the valve that is fixed by a seal holder 43 and a bracket 42. An axial SOW passage 44 that communicates between the inside of the seat member 34 accommodating the inertia valve 36 and the sealing liquid chamber C is formed in the seal holder final 41. The valve seat 38 is installed at a position slightly protruding from the inertia valve side opening of the liquid passage 44 so as to be able to engage with the inertia valve.
を九Δルツーディ２の弁座構成体背面側（図の右側）Ｋ
ｔ？轄る封止原素ＣＯ＠部に紘、弁座３６（シールホル
ダ）と同心をなすように外部と貫通する螺子孔４５を形
成し、この螺子孔４５に空気抜き穴４７を有する１リー
〆がルト４６を螺合せしめている。そしてこのツリ−〆
がルト４６は、図Ｏ右側からの螺出操作によ抄チー／臂
一部４８がパルｆ−ディ２の封止液室Ｃ内に形成した段
体弁座部４９に着座して、空気抜き穴４７を介しての封
止液室Ｃと外部との連通を遮断するが、反対に螺入操作
によりてはチーパ一部４８と段付弁座４９の着座が解除
されて封止液室Ｃを外部に連通させるよう設けられてい
ると共に、プリーダールト４６の螺子部とは反対側の端
部からは、シールホル〆３４の軸心部液通路４４内に遊
嵌嵌挿する小径の軸部５０が形成されており、この軸部
５００長さ社、プリーダ？ルト４６を螺入操作して封止
液室Ｃと外部を連通させたときにのみ、弁座３８から慣
性弁＊に突出することＫより慣性−パル３６と皺弁座３
８の当合を機械的に係止できるように設定されているの
である。The rear side of the valve seat structure of 9Δ Lutudy 2 (right side of the figure) K
T? A screw hole 45 that penetrates the outside is formed in the sealing element CO@ part that is concentric with the valve seat 36 (seal holder), and this screw hole 45 has a one-lead closure having an air vent hole 47. The bolt 46 is screwed together. Then, this tree-closing route 46 is screwed out from the right side in Figure O, so that the tip/arm part 48 is attached to the stepped valve seat portion 49 formed in the sealing liquid chamber C of the pallet f-day 2. When seated, the communication between the sealing liquid chamber C and the outside via the air vent hole 47 is cut off, but on the other hand, the seating of the tipper part 48 and the stepped valve seat 49 is released by the screwing operation. The sealing liquid chamber C is provided to communicate with the outside, and from the end of the pre-dart 46 on the opposite side from the threaded part, there is a small diameter hole that is loosely fitted into the axial liquid passage 44 of the seal holder 34. A shaft portion 50 is formed, and the length of this shaft portion 500 is made by Preda? Only when the bolt 46 is screwed in and the sealing liquid chamber C is communicated with the outside, the inertia valve * protrudes from the valve seat 38.
8 can be locked mechanically.
以上の液圧制御弁および慣性弁は、要するに、制御ピス
トン１７に作用する入・出力液ｊｉａｔｅａ１０液圧力
の差と、スゲリンダ機構からの付勢ΔネカのΔランスに
よりて、出力液室ａｌＯｌ［圧が入力ｆＫ ’１１ ａ
Ｓの液圧に対して所定Ｏ液圧折点値よｐ減圧制御され
ることＫ ｌｋ ｊｐ 、ζＯ液圧折点値は、封止液室
Ｃの液圧増大に応じ九スｌリンダ機構の制御ピストンへ
の付勢／母ネカの増大によって上昇されるのである。In short, the above-mentioned hydraulic pressure control valve and inertia valve control the output liquid chamber alOl [pressure is the input fK '11 a
The hydraulic pressure of S is controlled to be reduced by a predetermined O hydraulic pressure corner value. It is raised by increasing the bias/mother force on the control piston.
そして本発明の特徴は、とのスゲリンダ機構の構成、お
よびその作動に基づく作用・効果に特徴があり、これに
一ついて説明すると制御ピストン１７のスゲリング収容
部７に突出した他端部にはキヤ、ｆ状の第１のスゲリン
グシート２５が取着され、を九対向する調整ピストン３
２には第３のスゲリングシート２７が取着され、これら
第１・第３スプリングシー）２！Ｓ、２７０間に主たる
／４ネカを作用する九めの中スゲリンダ２９が張設され
ており、図示からも明らかなように１この中スデリンダ
２９によりて、制御ピストン１７は液圧制御弁の出力原
電ａｓ （図の左方）方向に押圧される。The present invention is characterized by the configuration of the sgellinder mechanism and the actions and effects based on its operation.One of the features of the present invention is the configuration of the sgellinder mechanism, and the actions and effects based on its operation. , an f-shaped first sgelling seat 25 is attached, and nine opposing adjusting pistons 3
2 is attached with a third spring seat 27, and these first and third spring seats) 2! A ninth snail cylinder 29 that acts as a main force is installed between S and 270, and as is clear from the diagram, the control piston 17 is controlled by the snail cylinder 29 to control the output of the hydraulic control valve. The original electric current is pressed in the direction (to the left in the figure).
塘九第１スゾリンダＶ−）２１５に対しては、これを調
整ピストン３２方向に押圧（すなわち中スｆｖンダ２９
を圧縮する方向に抑圧）しうるように第２０スゲリンダ
シート２６が配設されてお９、この中スゲリングシート
２６には、第１ｘｆリングシー）２８との間の内スｌリ
ンダ２８、パルプがディｌとの間の外スゲリンダ３０が
夫々張設され、これら内および外スゲリンダ２８．３０
によって前記中スプリング２９による制御ピストン１７
への付勢バネカを調整するように設けられている。なお
第２スプリングシート２６は第１ｘｆす／ダシ−トス５
と共に第３スプリングシート２７の方向に一定量移動す
ると、骸第３スゲリングシート２７と係合し、この後社
第１スプリングシート２５のみが第３スプリングシート
２７の方向に単独で移動しうるように設けられている。For the Togaku No. 1 Suzolinda V-) 215, press it in the direction of the adjustment piston 32 (i.e., press it in the direction of the adjustment piston 32
A 20th sliding cylinder seat 26 is disposed so as to be able to compress (in a compressing direction) 9, and this inner sliding cylinder 26 is provided with an inner sliding cylinder 28, An outer sgelinda 30 is provided between the pulp and the dil, respectively, and these inner and outer sgelindas 28.30
The control piston 17 by the middle spring 29
A biasing spring is provided to adjust the biasing spring. Note that the second spring seat 26 is the first xf seat/dash seat 5.
When the first spring seat 25 moves a certain amount in the direction of the third spring seat 27, it engages with the third spring seat 27, and after that, only the first spring seat 25 can move independently in the direction of the third spring seat 27. It is set in.
以上の組合せスゲリング機構の構成を要約すると、図示
する状態においては、第１および第２スプリングシー）
２１ｓ、２６が係合していることによりて、制御ピスト
ン１７に付勢されるバネ力社中スゲリンダ２９（初期バ
ネカｆ中、Ａネ定数に中とする）と外スデリンダ３０（
初期Δネカー、ｄネ定数咄とする）の関係でＦ中−Ｗ外
であり、封止液１１ｃ４り１ｍ圧が高くなることにより
て調１１−ストン３２が図の左方に移動畜れたと勤Ｏパ
ネカＯ増大によって、制御ピストン１７に作用するバネ
力は次のようになる。To summarize the configuration of the above combination sliding mechanism, in the illustrated state, the first and second spring seams are
21s and 26 are engaged, the control piston 17 is biased by the spring force inside the inner slider cylinder 29 (assuming that the initial spring force is in the middle and the A spring constant) and the outer slider cylinder 30 (
It is outside F-W due to the relationship between the initial Δneker and dne constant, and as the pressure of the sealing fluid 11c4 increases by 1m, the key 11-stone 32 moves to the left in the figure. As the force O is increased, the spring force acting on the control piston 17 becomes as follows.
ｒ中−Ｆ外＋ ｋ中Ｘ
したがって、この状態で制御ピストン１７が入・出力原
車ａｌｅ＠ｌの液圧作用により移動してこれら人・出力
液癩畠１＃ａｌの連通開閉を行なう場合は、更に（ｈｌ
ｌＰ−ｋ、） ｌ ｘ （ただし）Ｘは制御ピストン１
７の移動量）の／４ネカ変化を生ずるととになり、結局
Ｗ、ｐ −ｒ外＋ ｋ＊Ｘ ＋ （ｋ中 −ｋ
外） ノ 、 ・（−ｆ）が、液圧折点値を決定
するための付勢バネ力の値となる。r inside - outside F + inside k furthermore (hl
lP-k,) l x (However) X is control piston 1
7 movement amount), it becomes W, p −r outside + k*X + (k inside −k
(external) ノ, ・(-f) is the value of the biasing spring force for determining the hydraulic pressure corner value.
壕九車両の荷重積載量が大きいが故に封止液室ＣＫｊｌ
ＩＩｉい液圧が封止される場合には、前述の如く第２お
よび第３スゲリングシート２１ｉ、２？が係合し、他方
第１および第２スゲリングシート２５゜２６は離れした
が９てこの場合の減圧制御開始前における制御ピストン
１７に作用するバネ力は、Ｆ中 −ｒ内 ＋ ｌｃｔ、
１
（九だし１は少なくとも第２および第３スゲリングシー
）２６．２７が係合するに至る量）となる。Due to the large load carrying capacity of trench nine vehicles, the sealing liquid chamber CKjl
When a high hydraulic pressure is to be sealed, the second and third sliding sheets 21i, 2? The spring force acting on the control piston 17 before the start of pressure reduction control in the case where the first and second sgelling seats 25 and 26 are engaged and the first and second sgelling seats 25 and 26 are separated is F -r + lct,
1 (9 out of 1 is at least the second and third Sgelling Sea) 26.27 is the amount that results in engagement).
セしてこの後制御ピストン１７が移動したときのバネ力
変化によりて、液圧折点値を決定する丸めの付勢バネ力
の値は
Ｆ、ｔ、−ＦＡ ＋ ｋ中１ ＋（ｋ４．−に内）
ノｘ ＝（→（ただし内スプリング２８は初期バ
ネカＦ、Ｉバネ定数に内とする）
となる。The value of the rounded biasing spring force that determines the hydraulic pressure corner value is determined by the change in spring force when the control piston 17 moves after setting the pressure point. - inside)
No x = (→ (However, the inner spring 28 is assumed to have an initial spring force F and I spring constant).
而して前記式〇）、（→の対比より明らかなように、第
２および第３ス／リンダシー）！１１．２７０係合の前
後において、主たる／寸ネカを生ずる中スゲリンダ２９
の／ｆネカＦやに抗するＷＩｌ＊Ａネカの作用が外→内
のスｆ９ンダに変わる九めに、内スプリング２８のベネ
カＦ内を外スゲリンダｓＯのノダネカｒ外よりも小なる
ように設定すれば、
←）式 〉 （イ）式
となり、液圧折点値の増大割合はその分大なるもＯとな
る。Therefore, the above formula 〇), (as is clear from the comparison of →, the second and third stages/Lindasy)! 11.270 Medium sugelinda that causes major/dimensional inconsistency before and after engagement 29
In the 9th place, the action of WIl*A which resists /fnekaF changes from outside to inside sf9nda. If set, ←) formula 〉 (a) formula, and the increase rate of the hydraulic pressure breakpoint value becomes O.
このような構成のための各スゲリンダの初期ノ母ネカ、
バネ定数の関係は、本発明装置を適用する車両の制御特
性に応じて定められればよいが、原則的には次表の関係
を持つことがよい。The initial mother of each Sgelinda for such a composition,
The relationship between the spring constants may be determined depending on the control characteristics of the vehicle to which the device of the present invention is applied, but in principle it is preferable to have the relationship shown in the following table.
以上述べ九如く、本発明よりなる車両ブレーキ系に用い
る減速度感知式の液圧制御装置社、比較的簡単なる構成
によりて、１１Ｍｊツレー中力配分比により近似したグ
レー中液圧制御特性を得ることができ、しかも封止液室
に流入させる必要のある液量もスゲリンダ機構の圧縮変
位置が比較的小なる量で広範囲の付勢バネ力の変化管得
ることができるなどそのＭ層性は極めて大なるものであ
る。As described above, the deceleration sensing type hydraulic pressure control device used in a vehicle brake system according to the present invention obtains gray medium hydraulic pressure control characteristics that are similar to the 11Mj tree medium force distribution ratio with a relatively simple configuration. Moreover, the M-layer property is such that the amount of liquid that needs to flow into the sealing liquid chamber can be changed over a wide range of biasing spring force with a relatively small amount of compression displacement of the Sgelinda mechanism. It is extremely large.
図面社本発明の一実施例を示す液圧制御装置の縦断面図
である・
１．２−・・ベル！−ディ、 ３．４，５．６・・
・段付シリンダ１７・・・スゲリンダ収容部、 ８
−７．イルセイフピストン、９・・・Δランスシリンダ
、 １Ｇ−バックアップ、１１・−中シリンメ部材
、 １２−中シリンダ、１ Ｂ −・・バランスぜ
スＦン、 １４−・バルブシート、１５・・・ホール
トスｌリング、１ｇ−・・係止リング、１７−・・制御
ピストン、 １ｇ−・・７エイル七イフクリツプ、１
９・・・バルブシート、 ２Ｇ−ホールトスゲリ
ンダ、２１．２２・・・スプリング座、 ２３−ぜヌ
Ｆンカップ、２４−・係止りｙ／、 ２ ！！ ・
・１Ｈｘｆすｙｒｙ−ト、２＠−＠２スｌリンダシート
、２１−・第３スｌリンダシート、２８・・・内スゲリ
ンダ、 ２９・・・中スゲリンダ、３０・・・外スｌり
ンｒ１３１・・・調整シリン１１３２−・・調整ピスト
ン、 ３３−・慣性弁収容部、３４・・・シート部材、
３器−・ガイド面、３６・・・慣性？−ル、 ３
７・・・弁座構成体、３ ｇ −・・弁座、
３９・・・通液路、４０・・・通液開口 ４１−
・・通液開口、４２・−ブラダ、 ４３・・・
シールホルダ、４４−液通路、 ４５−・螺子孔
、４６・−１リーダールト、 ４７−・・空気抜き穴、
４８−−・チー／ｆ一部、 ４ Ｇ −・・段付弁座
、ｓＯ・・・軸部。Fig. 1 is a vertical cross-sectional view of a hydraulic pressure control device showing an embodiment of the present invention. 1.2-...Bell! -D, 3.4, 5.6...
・Stepped cylinder 17... Sgelinda housing part, 8
-7. Il-safe piston, 9...Δ lance cylinder, 1G-backup, 11-medium cylindrical member, 12-medium cylinder, 1B--balance valve seat, 14--valve seat, 15--hole toss l ring, 1g--locking ring, 17--control piston, 1g--7ail 7-if clip, 1
9...Valve seat, 2G-Holds Gelinda, 21.22...Spring seat, 23-Zenu F cup, 24-Latching y/, 2! !・
・1Hxf syry-to, 2@-@2 slinder seat, 21-・3rd slinder seat, 28...inner slider, 29...middle slider, 30...outer slider r131...Adjustment cylinder 1132--Adjustment piston, 33--Inertia valve housing portion, 34--Seat member,
3 instruments - guide surface, 36...inertia? -Le, 3
7... Valve seat structure, 3 g -... Valve seat,
39...Liquid passage, 40...Liquid passage opening 41-
・・Liquid passage opening, 42・-bladder, 43...
Seal holder, 44-liquid passage, 45-screw hole, 46-1 leader bolt, 47-air vent hole,
48--Chi/f part, 4 G--Stepped valve seat, sO... Shaft part.
Claims (1)
Ｋ伝える径路に介設され、液圧作用と付勢” ＊ 力の
バランスにて後輪ブレーキ液圧を折点減圧制御するデロ
ポーシ、二ング屋液圧制御弁と、車両制動時の減速度が
一定値を越えたときのブレーキ液圧管封止液富に封止す
るＧパルプと１備え、仁の封止液圧値に対し略比例的に
バネ力を増大するスプリング機構を介して前記液圧制御
弁への付勢ノ４ネカを増大させ液圧折点値を増大させる
形式の車両！レーキ系の減速度感知式液圧制御装置にお
いて、前記スプリング機構は、液圧制御弁にバネ力を付
勢する主たるスゲリングと、この主たるスプリングに抗
したバネ力を作用しうる２本の調整用スプリングとｔＮ
Ｌ、前記封止液室の液圧を受けて移動する１５に設けら
れ九調整ピストンの移動量が一定値を越えるときの前後
で、前記２本の調整用スゲリングの一々ネカ作用の保合
が切換わりて増大するよう構成したことを特徴とする車
両ブレーキ系の減速度感知式液圧制御装置。It is interposed in the path that transmits brake fluid pressure from the rear wheel brake system K from the brake fluid pressure cylinder to the rear wheel brake system K, and controls the rear wheel brake fluid pressure at a turning point by balancing the hydraulic pressure and energizing force. Equipped with a hydraulic pressure control valve and a G pulp that seals the brake fluid pressure pipe sealing fluid when the deceleration during vehicle braking exceeds a certain value, and is approximately proportional to the sealing fluid pressure value. In a vehicle! rake system deceleration sensing type hydraulic pressure control device of the type that increases the hydraulic pressure corner value by increasing the bias force to the hydraulic pressure control valve through a spring mechanism that increases the spring force, The spring mechanism includes a main spring ring that applies a spring force to the hydraulic control valve, two adjustment springs that can apply a spring force against this main spring, and tN.
L, the adjustment piston provided at 15 which moves in response to the liquid pressure in the sealing liquid chamber ensures that the negative action of each of the two adjustment pistons is maintained before and after the amount of movement of the adjustment piston exceeds a certain value; A deceleration sensing type hydraulic pressure control device for a vehicle brake system, characterized in that the device is configured to switch and increase the hydraulic pressure.
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP10431481A JPS588453A (en) | 1981-07-03 | 1981-07-03 | Deceleration sensing-fluid pressure controller of vehicle brake system |
Applications Claiming Priority (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP10431481A JPS588453A (en) | 1981-07-03 | 1981-07-03 | Deceleration sensing-fluid pressure controller of vehicle brake system |
Publications (1)
| Publication Number | Publication Date |
|---|---|
| JPS588453A true JPS588453A (en) | 1983-01-18 |
Family
ID=14377462
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| JP10431481A Pending JPS588453A (en) | 1981-07-03 | 1981-07-03 | Deceleration sensing-fluid pressure controller of vehicle brake system |
Country Status (1)
| Country | Link |
|---|---|
| JP (1) | JPS588453A (en) |
Cited By (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JP2015063154A (en) * | 2013-09-24 | 2015-04-09 | Ntn株式会社 | Electric brake device system |
Citations (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPS55119548A (en) * | 1979-03-09 | 1980-09-13 | Nissan Motor Co Ltd | Liquid pressure control valve |
-
1981
- 1981-07-03 JP JP10431481A patent/JPS588453A/en active Pending
Patent Citations (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPS55119548A (en) * | 1979-03-09 | 1980-09-13 | Nissan Motor Co Ltd | Liquid pressure control valve |
Cited By (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JP2015063154A (en) * | 2013-09-24 | 2015-04-09 | Ntn株式会社 | Electric brake device system |
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