JPH03224831A - Power transmission device for four-wheel drive vehicle - Google Patents
Power transmission device for four-wheel drive vehicleInfo
- Publication number
- JPH03224831A JPH03224831A JP23824590A JP23824590A JPH03224831A JP H03224831 A JPH03224831 A JP H03224831A JP 23824590 A JP23824590 A JP 23824590A JP 23824590 A JP23824590 A JP 23824590A JP H03224831 A JPH03224831 A JP H03224831A
- Authority
- JP
- Japan
- Prior art keywords
- torque
- fluid pressure
- rear wheels
- clutch
- wheels
- 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.)
- Granted
Links
- 230000005540 biological transmission Effects 0.000 title claims abstract description 59
- 239000012530 fluid Substances 0.000 abstract description 69
- 230000007423 decrease Effects 0.000 description 8
- 230000001133 acceleration Effects 0.000 description 7
- 238000010586 diagram Methods 0.000 description 5
- 230000000694 effects Effects 0.000 description 3
- 238000004891 communication Methods 0.000 description 2
- 230000008878 coupling Effects 0.000 description 2
- 238000010168 coupling process Methods 0.000 description 2
- 238000005859 coupling reaction Methods 0.000 description 2
- 230000006378 damage Effects 0.000 description 2
- 239000002184 metal Substances 0.000 description 2
- 230000002441 reversible effect Effects 0.000 description 2
- 239000013585 weight reducing agent Substances 0.000 description 2
- 239000003795 chemical substances by application Substances 0.000 description 1
- 230000001010 compromised effect Effects 0.000 description 1
- 238000005516 engineering process Methods 0.000 description 1
- 239000007769 metal material Substances 0.000 description 1
- 238000011017 operating method Methods 0.000 description 1
- 230000001737 promoting effect Effects 0.000 description 1
- 238000010992 reflux Methods 0.000 description 1
- 230000001052 transient effect Effects 0.000 description 1
- 238000011144 upstream manufacturing Methods 0.000 description 1
Landscapes
- Arrangement And Driving Of Transmission Devices (AREA)
Abstract
Description
【発明の詳細な説明】
［発明の目的］
〈産業上の利用分野〉
本発明は、前輪と後輪とを共通のエンジンにて駆動し得
るように構成された４輪駆動車輌の動力伝達装置に関す
る。[Detailed Description of the Invention] [Object of the Invention] <Industrial Application Field> The present invention provides a power transmission device for a four-wheel drive vehicle configured to drive front wheels and rear wheels with a common engine. Regarding.
〈従来の技術〉
４輪駆動車輌の一型式として、前・後車軸の一方をエン
ジンに直接的に連結し、この一方の車軸（主駆動軸）か
ら、相対回転速度応動型の粘性流体継手を介して他方の
車軸（従駆動軸）へと駆動トルクを伝達するようにした
ものが知られている。<Prior art> As a type of four-wheel drive vehicle, one of the front and rear axles is directly connected to the engine, and a relative rotational speed responsive type viscous fluid coupling is connected from this one axle (main drive shaft). It is known that drive torque is transmitted to the other axle (sub-drive shaft) via the drive shaft.
このような粘性流体継手は、主・従駆動軸間の回転速度
差に応じて伝達トルクが変化する特性を有しており、主
・従駆動軸間の回転速度差がある限度を超えると主・従
側駆動軸が略直結状態となる。Such viscous fluid couplings have the characteristic that the transmitted torque changes depending on the rotation speed difference between the main and slave drive shafts, and when the rotation speed difference between the main and slave drive shafts exceeds a certain limit, the main・The slave drive shaft is almost directly connected.
従って、この点について見ると、主・従側駆動軸の耐ト
ルク強度を同等に設定する必要がある。その一方、主・
従駆動軸間の回転速度差が極めて小さい状態にあっては
、従駆動軸への伝達トルクは実質的に０に等しく、この
状態における従駆動軸の負担は極めて軽い。このような
事情に鑑みて、従駆動軸への伝達トルクの上限を規定す
ることにより、従駆動軸側部材の実質的な負担を軽減し
、駆動系全体としての軽量化を企図しようとする技術が
、特開昭６３−４９５２６号公報に提案されている。Therefore, in this regard, it is necessary to set the torque resistance strengths of the main and slave drive shafts to be the same. On the other hand, the main
When the rotation speed difference between the slave drive shafts is extremely small, the torque transmitted to the slave drive shafts is substantially equal to 0, and the load on the slave drive shafts in this state is extremely light. In view of these circumstances, this technology aims to reduce the weight of the drive system as a whole by reducing the substantial burden on the slave drive shaft side members by stipulating the upper limit of the torque transmitted to the slave drive shaft. is proposed in Japanese Patent Application Laid-Open No. 63-49526.
〈発明が解決しようとする課題〉
ところで、予備タイヤを装着したり、あるいは積雪路走
行時に滑り止めを装着するなどして、前・後両輪の実質
的な有効径が互いに異なる状態での連続走行を余儀なく
されることがある。このような状態においては、常時前
・後輪間に回転速度差が生じ、本来は主・従側駆動軸間
にてトルク伝達を行なう必要のない走行状態である場合
にも、従駆動軸に必要以上に大きなトルクが伝達される
ことになる。つまり、通常走行にあっては、主・従駆動
軸間のトルク伝達は過渡状態のみを考慮すれば足りるの
に対し、上記のようにタイヤ径が異なる場合には、従駆
動軸に対して連続的にトルクが伝達されることになる。<Problems to be Solved by the Invention> By the way, continuous driving with the front and rear wheels having different effective diameters, such as by installing a spare tire or installing anti-slip devices when driving on a snowy road, etc. You may be forced to do so. In such conditions, there is always a difference in rotational speed between the front and rear wheels, and even in driving conditions where there is no need to transmit torque between the main and slave drive axles, there is a constant difference in rotational speed between the front and rear wheels. A larger torque than necessary will be transmitted. In other words, during normal driving, it is sufficient to consider only the transient state of torque transmission between the main and slave drive axles, whereas when the tire diameters differ as mentioned above, it is necessary to consider the torque transmission between the main and slave drive axles continuously. Torque will be transmitted accordingly.
一般に金属材料の機械的性質として、破壊応力以下の応
力であっても、これが繰返し作用すると、所謂金属疲労
を生じて破壊応力以下の応力によっても破壊に至ること
が知られている。特に高速走行状態にあっては、単位時
間当たりに加えられる応力の繰返し回数が多くなるので
、より一層疲労が進行し易くなり、限界応力の実質的な
低下を招くことが考えられる。従って、上記のような主
・従側駆動輪の有効径が互いに異なる状態での連続走行
にも耐え得るようにするには、その分室全率を高く設定
せねばならなくなり、前記した従来技術のように、従駆
動軸への伝達トルクの上限を規定するだけでは、軽量化
の達成が現実には十分になし得ないという不都合がある
。In general, it is known that as a mechanical property of metal materials, even if the stress is less than the breaking stress, if this stress is applied repeatedly, so-called metal fatigue will occur, leading to destruction even if the stress is less than the breaking stress. Particularly in high-speed running conditions, the number of repetitions of stress applied per unit time increases, so fatigue progresses even more easily, leading to a substantial decrease in critical stress. Therefore, in order to withstand continuous running in a state where the effective diameters of the main and slave drive wheels are different from each other as described above, it is necessary to set the total ratio of the compartments to a high value. As such, there is a problem in that it is not possible to achieve sufficient weight reduction in reality just by specifying the upper limit of the torque transmitted to the slave drive shaft.
本発明は、このような不都合を解消すべく案出されたも
のであり、その主な目的は、主・従駆動輪の径が互いに
異なる状態で走行することを考慮した上で、金属疲労を
誘発するような継続的な負荷トルクが従駆動側部材に対
して作用することのないように改善された４輪駆動車輌
の動力伝達装置を提供することにある。The present invention was devised to eliminate such inconveniences, and its main purpose is to reduce metal fatigue by taking into consideration that the main and slave drive wheels are driven with different diameters. It is an object of the present invention to provide a power transmission device for a four-wheel drive vehicle that is improved so that a continuous load torque that may be induced does not act on a subordinate drive side member.
［発明の構成］
〈課題を解決するための手段〉
このような目的は、本発明によれば、前輪と連動回転す
る第１部材と、後輪と連動回転する第２部材と、前記第
１部材と前記第２部材との間に介設された前記前輪と前
記後輪との回転速度差に応じて伝達トルクが変化するト
ルク伝達装置とを有する４輪駆動車輌の動力伝達装置で
あって、前記トルク伝達装置の伝達トルクの上限を規定
する伝達トルク制限手段と、前記トルク伝達装置の伝達
トルクを車速の増大に応じて減じる手段とを有すること
を特徴とする４輪駆動車輌の動力伝達装置を提供するこ
とにより達成される。[Structure of the Invention] <Means for Solving the Problems> According to the present invention, the object is to provide a first member that rotates in conjunction with the front wheel, a second member that rotates in conjunction with the rear wheel, and a first member that rotates in conjunction with the front wheel. A power transmission device for a four-wheel drive vehicle, comprising: a torque transmission device interposed between a member and the second member, the torque transmission device changing the transmission torque according to a rotational speed difference between the front wheel and the rear wheel; , a power transmission for a four-wheel drive vehicle, comprising a transmission torque limiting means for defining an upper limit of the transmission torque of the torque transmission device, and a means for reducing the transmission torque of the torque transmission device according to an increase in vehicle speed. This is accomplished by providing a device.
く作用〉
このような構成によれば、主・従側駆動軸間の伝達トル
クの上限が適宜な所定値に規定され、かつ伝達トルク容
量が走行速度の増大と共に減少する。従って、特に主駆
動輪がスリップし易い発進加速時（低速時）には、主駆
動輪から従駆動輪へのトルク伝達が十分になされ、主駆
動輪がスリップする可能性が低い高速走行時には、実質
的な伝達トルクが減少する。従って、前輪と後輪との径
が互いに異なる（特に主駆動輪の径がより小さい）状態
での連続走行における従駆動軸側部材への駆動トルクの
伝達を軽減し得ることから、従駆動軸側部材の強度余裕
の設定を低減し得る。Effect> According to such a configuration, the upper limit of the transmission torque between the main and slave drive shafts is set to an appropriate predetermined value, and the transmission torque capacity decreases as the traveling speed increases. Therefore, during start acceleration (at low speeds) when the main drive wheels are likely to slip, sufficient torque is transmitted from the main drive wheels to the sub drive wheels, and during high speed driving when the possibility of the main drive wheels slipping is low. Substantial transmitted torque is reduced. Therefore, it is possible to reduce the transmission of drive torque to the slave drive shaft side member during continuous running when the diameters of the front wheels and rear wheels are different from each other (especially when the diameter of the main drive wheel is smaller). The setting of the strength margin of the side member can be reduced.
〈実施例〉
以下、添付の図面を参照して本発明の好適実施例につい
て詳細に説明する。<Embodiments> Hereinafter, preferred embodiments of the present invention will be described in detail with reference to the accompanying drawings.
第１図は、本発明に基づく動力伝達装置が適用された４
輪駆動車輌の動力伝達系を示すスケルトン図である。エ
ンジン１の出力は、変速機２を介して前輪側の差動装置
３に入力する。そして差動装置３の出力は、ドライブシ
ャフト４を介して左右各前輪５に伝達される。FIG. 1 shows a four-wheel drive system to which the power transmission device according to the present invention is applied.
FIG. 2 is a skeleton diagram showing a power transmission system of a wheel drive vehicle. The output of the engine 1 is input to a front wheel differential device 3 via a transmission 2. The output of the differential device 3 is transmitted to the left and right front wheels 5 via the drive shaft 4.
差動装置３に入力したエンジン１の出力は、傘歯車装置
６を介して後記する動力伝達装置７に入力し、該動力伝
達装置７の出力は、傘歯車装置８を介して後輪側の差動
装置９に伝達される。そして差動装置９の出力は、ドラ
イブシャフト１０を介して左右各後輪１１に伝達される
。The output of the engine 1 that is input to the differential device 3 is input to a power transmission device 7 (described later) via a bevel gear device 6, and the output of the power transmission device 7 is transmitted to the rear wheel side via a bevel gear device 8. The signal is transmitted to the differential gear 9. The output of the differential gear 9 is transmitted to the left and right rear wheels 11 via the drive shaft 10.
動力伝達装置７は、前輪側の傘歯車装置６の出力軸に連
動駆動される第１流体圧ポンプ２１と、後輪側の傘歯車
装置８の入力軸に連動駆動される第２流体圧ポンプ２２
と、前輪側傘歯車装置６の出力軸と後輪側傘歯車装置８
の入力軸との間に介設されたトルク伝達装置としての流
体圧作動クラッチ２３と、第１・第２両流体圧ポンプ２
１・２２及びクラッチ２３に係わるオイルの流れを制御
する流体圧制御回路（後に詳述）とからなっている。The power transmission device 7 includes a first fluid pressure pump 21 that is driven in conjunction with the output shaft of the bevel gear device 6 on the front wheel side, and a second fluid pressure pump that is driven in conjunction with the input shaft of the bevel gear device 8 on the rear wheel side. 22
, the output shaft of the front wheel side bevel gear device 6 and the rear wheel side bevel gear device 8
A fluid pressure operated clutch 23 as a torque transmission device interposed between the input shaft of the first and second fluid pressure pumps 2 and the first and second fluid pressure pumps 2
1 and 22 and a fluid pressure control circuit (described in detail later) that controls the flow of oil related to the clutch 23.
ここで前・後側傘歯車装置６・８のギヤ比が互いに異な
る値にされており、流体圧ポンプ回転速度と車輪回転速
度との関係は、次に示す関係になっている。Here, the gear ratios of the front and rear bevel gear devices 6 and 8 are set to different values, and the relationship between the fluid pressure pump rotation speed and the wheel rotation speed is as shown below.
前輪回転速度 後輪回転速度これはすなわち
、車輪回転速度に対するポンプ回転速度の増大率が第２
流体圧ポンプ２２の方がより大きいことを意味しており
、第３図に示すように、前・後両輪５・１１の回転速度
が同一であれば、第２流体圧ポンプ２２の回転速度がよ
り高くなるようになっており、しかも両流体圧ポンプ２
１・２２の回転速度差は、車輪回転速度に正比例して増
大することを意味している。Front wheel rotation speed Rear wheel rotation speed This means that the increase rate of the pump rotation speed with respect to the wheel rotation speed is the second
This means that the fluid pressure pump 22 is larger, and as shown in FIG. 3, if the rotation speeds of the front and rear wheels 5 and 11 are the same, the rotation speed of the second fluid pressure pump 22 is It is designed to be higher, and both hydraulic pumps 2
A rotational speed difference of 1.22 is meant to increase in direct proportion to the wheel rotational speed.
第１流体圧ポンプ２１は、ギヤポンプあるいはベーンポ
ンプからなり、車輌が前進時には吐出ポートとなり後退
時には吸入ポートとなる第１ポート２４と、前進時には
吸入ポートとなり後退時には吐出ポートとなる第２ポー
ト２５とを有している。そして第２流体圧ポンプ２２は
、同じくギヤポンプあるいはベーンポンプからなり、車
輌が前進時には吸入ポートとなり後退時には吐出ポート
となる第３ポート２６と、前進時には吐出ポートとなり
後退時には吸入ポートとなる第４ポート２７とを有して
いる。これら各ポート２４〜２７は、第１ポート２４と
第３ポート２６とが第１連結油路２８を介して連通接続
され、第２ポート２５と第４ポート２７とが第２連結油
路２９を介して連通接続されている。The first fluid pressure pump 21 is a gear pump or a vane pump, and has a first port 24 that becomes a discharge port when the vehicle is moving forward and an intake port when it is reversing, and a second port 25 that becomes an intake port when the vehicle is moving forward and a discharge port when it is reversing. have. The second fluid pressure pump 22 is also a gear pump or a vane pump, and has a third port 26 that becomes an intake port when the vehicle is moving forward and a discharge port when it is reversing, and a fourth port 27 that becomes a discharge port when the vehicle is moving forward and an intake port when it is reversing. It has In each of these ports 24 to 27, the first port 24 and the third port 26 are connected to each other via the first connecting oil passage 28, and the second port 25 and the fourth port 27 are connected to each other through the second connecting oil passage 29. It is connected through.
ここで第１・第２両流体圧ポンプ２１・２２は、そのチ
ャンバ容積が互いに異なっており、第１流体圧ポンプ２
１の一回転光たりの吐出量が、第２流体圧ポンプ２２の
一回転光たりの吐出量に比してより小さい設定になって
いる。これは前・後車軸４・１０とポンプ軸との速比が
同一であり、かつ前・後輪５・１１の有効径が同一であ
れば、車輪回転速度に応じた吐出量の変化率は、第２流
体圧ポンプ２２の方がより大きく、両流体圧ポンプ２１
・２２の吐出量差は、車輪回転速度に正比例して増大す
ることを意味している（第３図）。Here, both the first and second fluid pressure pumps 21 and 22 have different chamber volumes, and the first and second fluid pressure pumps 21 and 22 have different chamber volumes.
The discharge amount per one revolution of the second fluid pressure pump 22 is set to be smaller than the discharge amount per one revolution of the second fluid pressure pump 22 . This means that if the speed ratios of the front and rear axles 4 and 10 and the pump shaft are the same, and the effective diameters of the front and rear wheels 5 and 11 are the same, the rate of change in the discharge amount according to the wheel rotation speed is , the second fluid pressure pump 22 is larger, and both fluid pressure pumps 21
- The discharge amount difference of 22 means that it increases in direct proportion to the wheel rotation speed (Fig. 3).
第１連結油路２８及び第２連結油路２９と、流体圧作動
クラッチ２３の作動油圧室３０との間は、切換弁３１を
介して連結されている。この切換弁３１は、主に変速機
２が前進段にあるか、あるいは後退段にあるかに応じて
切換わるスプール弁からなり、２つに仕切られた弁室３
２・３３と、第１弁室３２から第２弁室３３へ向けての
流れを規制する一方向弁３４と、第１弁室３２と第２室
弁室３３との差圧が所定値になると、第１弁室３２と第
２弁室３３との間を連通し、第１弁室３２から第２弁室
３３へ向けての流れを許容するリリーフ弁３５とを有し
ている。この切換弁３１の作動により、前進時にあって
は、第１図に示すように、第２連結油路２９とオイルタ
ンク３６との間が第２弁室３３を介して連通し、第１連
結油路２８とクラッチの作動油圧室３０との間が、バイ
パス油路３７ａ・第１弁室３２・作動油圧供給油路３７
ｂを介して連通し、しかもクラッチの作動油圧室３０に
作用する圧力が所定値以上になると、リリーフ弁３５を
介してオイルタンク３６へ圧力が逃げるようになってい
る。そして後退時にあっては、第２図に示すように、第
１連結油路２８とオイルタンク３６との間が第２弁室３
３を介して連通し、第２連結油路２９とクラッチの作動
油圧室３０との間が第１弁室３２を介して連通し、しか
もクラッチの作動油圧室３０に作用する圧力が所定値以
上になると、リリーフ弁３５を介してタンク３６へ圧力
が逃げるようになっている。The first connecting oil passage 28 and the second connecting oil passage 29 are connected to the hydraulic pressure chamber 30 of the hydraulically operated clutch 23 via a switching valve 31. This switching valve 31 mainly consists of a spool valve that switches depending on whether the transmission 2 is in forward gear or reverse gear, and has a valve chamber 3 partitioned into two.
2.33, a one-way valve 34 that regulates the flow from the first valve chamber 32 to the second valve chamber 33, and a pressure difference between the first valve chamber 32 and the second valve chamber 33 to a predetermined value. Then, it has a relief valve 35 that communicates between the first valve chamber 32 and the second valve chamber 33 and allows flow from the first valve chamber 32 to the second valve chamber 33. Due to the operation of this switching valve 31, during forward movement, as shown in FIG. A bypass oil passage 37a, a first valve chamber 32, and an operating oil pressure supply oil passage 37 are connected between the oil passage 28 and the clutch working oil pressure chamber 30.
When the pressure acting on the hydraulic pressure chamber 30 of the clutch reaches a predetermined value or more, the pressure escapes to the oil tank 36 via the relief valve 35. When reversing, as shown in FIG. 2, the space between the first connecting oil passage 28 and the oil tank 36 is
3, the second connecting oil passage 29 and the clutch working hydraulic chamber 30 communicate with each other through the first valve chamber 32, and the pressure acting on the clutch working hydraulic chamber 30 is equal to or higher than a predetermined value. When this happens, pressure escapes to the tank 36 via the relief valve 35.
更に、第１弁室３２とクラッチの作動油圧室３０との間
を連結する作動油圧供給油路３７ｂは、オリフィス３８
を有する分岐通路を介してタンク３６の油面上に連通し
ている。Furthermore, the working hydraulic pressure supply oil passage 37b connecting the first valve chamber 32 and the working hydraulic pressure chamber 30 of the clutch has an orifice 38.
It communicates with the oil surface of the tank 36 through a branch passage having a diameter.
０
次に上記実施例の作動の要領について各状態に応じて順
に説明する。0 Next, the operation of the above embodiment will be explained in order according to each state.
前進発進加速時には、後輪１１が停止したままで前輪５
のみがスリップ状態で回転することがある。この時には
、前輪５と共に第１流体圧ポンプ２１のみが回転するた
め、オイルタンク３６から第２弁室３３及び第２連結油
路２９を介して第２ポート２５に吸入されたオイルは、
第１ポート２４から第１連結油路２８へ吐出されてバイ
パス油路３７ａに全量が流入し、第１弁室３２及び作動
油圧供給油路３７ｂを介してクラッチの作動油圧室３０
に油圧を作用させる。これによりクラッチ２３が係合し
、前輪５と後輪１１との間が連結される。During forward acceleration, the rear wheels 11 remain stationary and the front wheels 5
The chisel may rotate in a slip state. At this time, since only the first fluid pressure pump 21 rotates together with the front wheel 5, the oil sucked into the second port 25 from the oil tank 36 via the second valve chamber 33 and the second connecting oil passage 29,
The entire amount is discharged from the first port 24 to the first connecting oil passage 28 and flows into the bypass oil passage 37a, and then passes through the first valve chamber 32 and the hydraulic pressure supply passage 37b to the hydraulic pressure chamber 30 of the clutch.
apply hydraulic pressure to. As a result, the clutch 23 is engaged, and the front wheels 5 and the rear wheels 11 are connected.
ここでクラッチの伝達トルクは、オリフィス３８の流量
によって定まるオリフィス上流側の油圧に正比例し、こ
の油圧は、両流体圧ポンプ２１・２２の吐出量（吸入量
）差の２乗に比例して変化する。また、リリーフ弁３５
の開放圧の設定により、クラッチ２３の伝達トルクの上
限値を適宜に１
設定することができる（第４図）。Here, the transmission torque of the clutch is directly proportional to the oil pressure on the upstream side of the orifice determined by the flow rate of the orifice 38, and this oil pressure changes in proportion to the square of the difference in discharge amount (suction amount) between the two fluid pressure pumps 21 and 22. do. In addition, the relief valve 35
By setting the opening pressure of , the upper limit value of the transmission torque of the clutch 23 can be appropriately set to 1 (FIG. 4).
クラッチ２３が係合して後輪側に駆動トルクが分配され
ると、後輪１１の回転速度の増大に応じて第１流体圧ポ
ンプ２１の吐出油が第２流体圧ポンプ２２に吸入される
ようになる。そして第１流体圧ポンプ２１の吐出量と第
２流体圧ポンプ２２の吸入量との差に応じてクラッチ２
３の係合力、すなわち後輪への伝達トルクが自動的に変
化し、第１・第２両流体圧ポンプ２１・２２の吐出量（
吸入量）が互いにバランスすると、作動油圧供給油路３
７ｂへの吐出圧は発生しなくなり、クラッチ２３の係合
が断たれる。When the clutch 23 is engaged and drive torque is distributed to the rear wheels, oil discharged from the first fluid pressure pump 21 is sucked into the second fluid pressure pump 22 in accordance with an increase in the rotational speed of the rear wheels 11. It becomes like this. Then, the clutch 2
The engagement force of No. 3, that is, the torque transmitted to the rear wheels, changes automatically, and the discharge amount (
When the suction amount) is balanced with each other, the hydraulic pressure supply oil path 3
The discharge pressure to 7b is no longer generated, and the clutch 23 is disengaged.
ここでエンジン１に直接的に駆動される前輪５に連動駆
動される第１流体圧ポンプ２１と、動力伝達装置７を介
して駆動力を伝達される後輪１１に連動駆動される第２
流体圧ポンプ２２との吐出量（吸入量）のバランス点は
、両流体圧ポンプ２１・２２の運転特性が前記したよう
に第３図に示す関係に設定されていることから、前輪５
の回転速度が後輪１１よりも高い時点で現れる。そして
２
このバランス点での前・後両輪５・１１の回転速度差は
、車速が高くなるほど大きくなる。これらの特性は、前
・後両輪５・１１の同一の回転速度差に対するクラッチ
伝達トルクの大きさが、車速が高くなるほど減少するこ
とを示している。このことは、クラッチ２３の伝達トル
ク容量、すなわち差動制限力は、車速が高くなるほど減
少することを意味している（第４図）。Here, a first fluid pressure pump 21 is driven in conjunction with the front wheels 5 which are directly driven by the engine 1, and a second fluid pressure pump 21 is driven in conjunction with the rear wheels 11 to which driving force is transmitted via the power transmission device 7.
The balance point of the discharge amount (suction amount) with the fluid pressure pump 22 is determined by the fact that the operating characteristics of both the fluid pressure pumps 21 and 22 are set to the relationship shown in FIG.
appears when the rotational speed of the rear wheel 11 is higher than that of the rear wheel 11. And 2. The rotational speed difference between the front and rear wheels 5 and 11 at this balance point increases as the vehicle speed increases. These characteristics indicate that the magnitude of the clutch transmission torque for the same rotational speed difference between the front and rear wheels 5 and 11 decreases as the vehicle speed increases. This means that the transmission torque capacity of the clutch 23, that is, the differential limiting force, decreases as the vehicle speed increases (FIG. 4).
前進緩加速時、緩減速時及び定速走行時にあっては、前
輪５と後輪１１とが同一径であれば、両輪は略同−回転
速度で回転する。そして前・後両輪５・１１が同一回転
速度であれば、第１流体圧ポンプ２１の吐出量が第２流
体圧ポンプ２２の吸入量を常に下回り、かつ第２流体圧
ポンプ２２の吐出量が第１流体圧ポンプ２１の吸入量を
常に上回ることになる。すると第１ポート２４からの吐
出油は専ら第３ポート２６に吸入され、かつ第４ポート
２７からの吐出油の一部は第２連結油路２９・第２弁室
３３・一方向弁３４・第１弁室３２・バイパス油路３７
ａ・第１連結油路２８を経て３
第３ポート２６へ環流する。この結果、第１連結油路２
８の管内圧はクラッチ２３の作動圧に到達せず、後輪１
１に対して駆動力が伝達されない。During slow forward acceleration, slow deceleration, and constant speed running, if the front wheels 5 and rear wheels 11 have the same diameter, they rotate at approximately the same rotational speed. If both the front and rear wheels 5 and 11 have the same rotational speed, the discharge amount of the first fluid pressure pump 21 is always lower than the suction amount of the second fluid pressure pump 22, and the discharge amount of the second fluid pressure pump 22 is always lower than the suction amount of the second fluid pressure pump 22. This will always exceed the suction amount of the first fluid pressure pump 21. Then, the oil discharged from the first port 24 is sucked exclusively into the third port 26, and a part of the oil discharged from the fourth port 27 is sucked into the second connecting oil passage 29, second valve chamber 33, one-way valve 34, First valve chamber 32/bypass oil passage 37
a. Returns to the third port 26 via the first connecting oil passage 28. As a result, the first connecting oil passage 2
The pressure inside the pipe 8 does not reach the operating pressure of the clutch 23, and the rear wheel 1
No driving force is transmitted to 1.
定速走行時に前輪５のみが摩擦係数の低い路面を踏んだ
場合、あるいは急加速せんとした時には、前輪５が過渡
的にスリップ状態になることがある。If only the front wheels 5 step on a road surface with a low coefficient of friction while driving at a constant speed, or if the vehicle attempts to accelerate suddenly, the front wheels 5 may slip temporarily.
このような状態においては、第１ポート２４からの吐出
量が第３ポート２６への吸入量を上回るほど前輪５の回
転速度が後輪１１のそれを上回ると、第１流体圧ポンプ
２１の吐出油を第２流体圧ポンプ２２が吸入しきれなく
なるため、両流体圧ポンプ２１・２２の吐出量（吸入量
）の差に対応した油圧が第１連結油路２８に発生する。In such a state, when the rotational speed of the front wheel 5 exceeds that of the rear wheel 11 to the extent that the discharge amount from the first port 24 exceeds the suction amount to the third port 26, the discharge amount of the first fluid pressure pump 21 decreases. Since the second fluid pressure pump 22 is no longer able to suck in the oil, a hydraulic pressure corresponding to the difference in the discharge amount (suction amount) between the two fluid pressure pumps 21 and 22 is generated in the first connecting oil path 28 .
この油圧は、バイパス油路３７ａ・第１弁室３２・作動
油圧供給油路３７ｂを経てクラッチの作動油圧室３０に
導かれる。これによりクラッチ２３が係合し、後輪１１
に対して駆動トルクが分配される。そしてクラッチ２３
が係合して後輪側へ駆動トルクが分配されると、上記と
同様にして前後輪間の回転速度差に応じてクラッチ２３
の係合力、すなわち後４
輪画へ伝達されるトルクの大きさが自動的に変化するが
、この場合、回転速度差に対する伝達トルクの大きさは
、車速の増大と共に小さくなる。This oil pressure is guided to the working oil pressure chamber 30 of the clutch via the bypass oil passage 37a, the first valve chamber 32, and the working oil pressure supply oil passage 37b. As a result, the clutch 23 is engaged, and the rear wheel 11
Driving torque is distributed to. and clutch 23
When the clutch 23 is engaged and drive torque is distributed to the rear wheels, the clutch 23
The engagement force, that is, the magnitude of the torque transmitted to the rear four wheels automatically changes, but in this case, the magnitude of the transmitted torque relative to the rotational speed difference decreases as the vehicle speed increases.
車輪に制動力が作用すると、前後輪の制動力配分は一般
に前輪側がより高く設定されているので、急制動時など
では、後輪１１よりも前輪５が先にロックする。また、
定速走行からのエンジンブレーキは前輪５にのみ作用す
るので、この場合も過渡的には前輪５の回転速度が後輪
１１よりも低くなる。そして前輪５の回転速度が後輪１
１に比して低くなると、第１流体圧ポンプ２１の吐出量
が第２流体圧ポンプ２２の吸入量を下回るため、作動油
圧供給油路３７ｂへの吐出圧は発生せず、クラッチ２３
は係合しない。従って、前後輪間の連結は断たれる。こ
のとき、第４ポート２７からの吐出油の一部は、第２連
結油路２９・第２弁室３３・一方向弁３４・第１弁室３
２・バイパス油路３７ａ・第１連結油路２８を経て第３
ポート２６へ環流する。When braking force is applied to the wheels, the braking force distribution between the front and rear wheels is generally set higher on the front wheel side, so the front wheels 5 lock before the rear wheels 11 during sudden braking. Also,
Since the engine brake from constant speed running acts only on the front wheels 5, the rotational speed of the front wheels 5 is transiently lower than that of the rear wheels 11 in this case as well. And the rotation speed of front wheel 5 is rear wheel 1
1, the discharge amount of the first fluid pressure pump 21 is lower than the suction amount of the second fluid pressure pump 22, so no discharge pressure is generated to the hydraulic pressure supply oil passage 37b, and the clutch 23
is not engaged. Therefore, the connection between the front and rear wheels is severed. At this time, a part of the oil discharged from the fourth port 27 is transferred to the second connecting oil passage 29, the second valve chamber 33, the one-way valve 34, and the first valve chamber 3.
2. Bypass oil passage 37a and the third via the first connecting oil passage 28
Reflux to port 26.
前輪５が完全にロックすると、第１流体圧ポン５
プ２１が停止して第２流体圧ポンプ２２のみが回転する
。すると第４ポート２７から第２連結油路２９への吐出
油は、第２弁室３３・一方向弁３４・第１弁室３２・バ
イパス油路３７ａ・第１連結油路２８を経て第３ポート
２６へと全量が環流する。従って、この場合もクラッチ
２３は係合せず、前後輪間の連結は断たれる。When the front wheels 5 are completely locked, the first hydraulic pump 21 stops and only the second hydraulic pump 22 rotates. Then, the oil discharged from the fourth port 27 to the second connecting oil passage 29 passes through the second valve chamber 33, the one-way valve 34, the first valve chamber 32, the bypass oil passage 37a, and the first connecting oil passage 28, and then flows to the third connecting oil passage 28. The entire amount flows back to port 26. Therefore, in this case as well, the clutch 23 is not engaged, and the connection between the front and rear wheels is severed.
後退時には、第１・第２両流体圧ポンプ２１・２２の回
転方向が共に逆になり、吐出ポートと吸入ポートとの関
係が上記とは逆の関係になるが、基本的な作動原理は前
進時と同様にして行なわれる。When moving backward, the rotational directions of both the first and second fluid pressure pumps 21 and 22 are reversed, and the relationship between the discharge port and the suction port is reversed to that described above, but the basic operating principle is when moving backward. It is done in the same way as before.
後退発進加速時には、−時的に第１流体圧ポンプ２１の
みが回転する。すると第２図に示すように、オイルタン
ク３６から第２弁室３３・バイパス油路３７ａ・第１連
結油路２８を経て第１ポート２４に吸入されたオイルは
、第２ポート２５から第２連結油路２９へ吐出され、第
１弁室３２及び作動油圧供給油路３７ｂを経てクラッチ
の作動油圧室３０に油圧を作用させる。これによりクラ
６
ツチ２３が接続し、後輪１１に駆動トルクが分配される
。During the backward acceleration, only the first fluid pressure pump 21 rotates. Then, as shown in FIG. 2, the oil sucked into the first port 24 from the oil tank 36 via the second valve chamber 33, the bypass oil passage 37a, and the first connecting oil passage 28 is transferred from the second port 25 to the second The oil is discharged to the connecting oil passage 29, and acts on the hydraulic pressure chamber 30 of the clutch through the first valve chamber 32 and the hydraulic pressure supply oil passage 37b. As a result, the clutch 23 is connected and drive torque is distributed to the rear wheels 11.
そして前進時と同様に、後輪側の回転速度の増大に応じ
て第１流体圧ポンプ２１の吐出油の一部が第２流体圧ポ
ンプ２２に吸入されるようになり、この時の両流体圧ポ
ンプ２１・２２の吐出量（吸入量）差に応じてクラッチ
の作動油圧室３０に作用する油圧が変化して後輪へのト
ルク分配率が変化し、両流体圧ポンプ２１・２２の吐出
量（吸入量）が互いにバランスした状態になると、クラ
ッチの作動油圧室３０に油圧が作用しなくなって前後輪
間の接続が断たれる。Then, like when moving forward, a part of the oil discharged from the first fluid pressure pump 21 is sucked into the second fluid pressure pump 22 as the rotational speed of the rear wheels increases, and both fluids at this time The hydraulic pressure acting on the hydraulic pressure chamber 30 of the clutch changes according to the difference in the discharge amount (suction amount) of the pressure pumps 21 and 22, and the torque distribution ratio to the rear wheels changes, causing the discharge amount of both the fluid pressure pumps 21 and 22 to change. When the amounts (intake amounts) are in balance with each other, hydraulic pressure ceases to act on the hydraulic pressure chamber 30 of the clutch, and the connection between the front and rear wheels is severed.
後退緩加速時、緩減速時及び定速走行時にあっては、前
進時と同様に、第１流体圧ポンプ２１の吐出量が第２流
体圧ポンプ２２の吸入量を常に下回り、かつ第２流体圧
ポンプ２２の吐出量が第１流体圧ポンプ２１の吸入量を
常に上回ることになる。すると第２ポート２５からの吐
出油が第４ポート２７に吸入され、かつ第３ポート２６
からの吐出油の一部が第１連結油路２８・バイパス油路
７
３７ａ・第２弁室３３・一方向弁３４・第１弁室３２・
第２連結油路２９を経て第４ポート２７へ環流する。こ
の結果、第２連結油路２９の管内圧はクラッチ２３の作
動圧に到達せず、後輪１１に対して駆動力は伝達されな
い。During slow backward acceleration, slow deceleration, and constant speed running, the discharge amount of the first fluid pressure pump 21 is always lower than the suction amount of the second fluid pressure pump 22, and the second fluid The discharge amount of the pressure pump 22 will always exceed the suction amount of the first fluid pressure pump 21. Then, the oil discharged from the second port 25 is sucked into the fourth port 27, and the oil is sucked into the third port 26.
A portion of the oil discharged from the first connecting oil passage 28, bypass oil passage 7 37a, second valve chamber 33, one-way valve 34, first valve chamber 32,
It flows back to the fourth port 27 via the second connection oil passage 29. As a result, the internal pressure of the second connecting oil passage 29 does not reach the operating pressure of the clutch 23, and no driving force is transmitted to the rear wheels 11.
後退定速走行からの急加速などにより前輪５がスリップ
状態になり、第１流体圧ポンプ２１の吐出量が第２流体
圧ポンプ２２の吸入量を上回るほど前輪５の回転速度が
後輪１１のそれを上回ると、第１流体圧ポンプ２１の吐
出油を第２流体圧ポンプ２２が吸入しきれなくなるため
、両ポンプ２１・２２の吐出量（吸入量）の差に対応し
た油圧が第２連結油路２９に発生する。この油圧は、第
１弁室３２・作動油圧供給油路３７ｂを経てクラッチの
作動油圧室３０に導かれる。これによりクラッチ２３が
係合し、後輪１１に対して駆動トルクが分配される。The front wheels 5 enter a slip state due to sudden acceleration from constant speed traveling in reverse, etc., and the more the discharge amount of the first fluid pressure pump 21 exceeds the intake amount of the second fluid pressure pump 22, the more the rotation speed of the front wheels 5 increases. If it exceeds this, the second fluid pressure pump 22 will not be able to suck in the oil discharged from the first fluid pressure pump 21, so the oil pressure corresponding to the difference in the discharge amount (suction amount) of both pumps 21 and 22 will be applied to the second connection. This occurs in the oil passage 29. This oil pressure is guided to the working oil pressure chamber 30 of the clutch via the first valve chamber 32 and the working oil pressure supply oil passage 37b. As a result, the clutch 23 is engaged and drive torque is distributed to the rear wheels 11.
後退制動時には、第１流体圧ポンプ２１の回転速度が第
２流体圧ポンプ２２のそれを下回るため、作動油圧供給
油路３７ｂへの吐出圧は発生せず、８
クラッチ２３は係合しない。従って、前後輪間の連結は
断たれる。このとき、第３ポート２６からの第２流体圧
ポンプ２２の吐出油の一部は、第１連結油路２８・バイ
パス油路３７ａ・第２弁室３３・一方向弁３４・第１弁
室３２・第２連結油路２９を経て第４ポート２７へ環流
する。そして前輪５が完全にロックすると、第３ポート
２６からの吐出油は、第１連結油路２８・バイパス油路
３７ａ・第２弁室３３・一方向弁３４・第１弁室３２・
第２連結油路２９を経て第４ポート２７へ全量が環流す
る。従って、この場合もクラッチ２３は係合せず、前後
輪間の連結は断たれる。During backward braking, the rotation speed of the first fluid pressure pump 21 is lower than that of the second fluid pressure pump 22, so no discharge pressure is generated to the hydraulic pressure supply oil passage 37b, and the 8 clutch 23 is not engaged. Therefore, the connection between the front and rear wheels is severed. At this time, a part of the oil discharged from the second fluid pressure pump 22 from the third port 26 is transferred to the first connecting oil passage 28, the bypass oil passage 37a, the second valve chamber 33, the one-way valve 34, and the first valve chamber. 32 and flows back to the fourth port 27 via the second connecting oil passage 29. When the front wheel 5 is completely locked, the oil discharged from the third port 26 is transmitted to the first connecting oil passage 28, bypass oil passage 37a, second valve chamber 33, one-way valve 34, first valve chamber 32,
The entire amount flows back to the fourth port 27 via the second connecting oil passage 29. Therefore, in this case as well, the clutch 23 is not engaged, and the connection between the front and rear wheels is severed.
次に前・後輪の有効径が互いに異なる場合の動作につい
て説明する。Next, the operation when the front and rear wheels have different effective diameters will be explained.
エンジン１に直接的に駆動される前輪５の有効径が動力
伝達装置７を介して駆動される後輪１１の有効径よりも
小さい場合を想定すると、この場合には、前・後両輪５
・１１が共にスリップもロックもしていない定速走行状
態に達しても、前輪回転速度が後輪回転速度を常に」二
回ることになる。Assuming that the effective diameter of the front wheels 5 that are directly driven by the engine 1 is smaller than the effective diameter of the rear wheels 11 that are driven via the power transmission device 7, in this case, both the front and rear wheels 5
・Even if both wheels reach a constant speed running state with neither slipping nor locking, the front wheel rotational speed will always be twice the rear wheel rotational speed.
９
ここで前・後輪５・１１の有効径差は定数であるから、
走行速度に正比例して両輪５・１１の回転速度差が増大
する。すると第１・第２両流体圧ポンプ２１・２２の吐
出量（吸入量）の差に応じてクラッチ２３の係合力が高
まることになるが、吐出量（吸入量）の差が大きくなり
、クラッチの作動油圧室３０に作用する油圧が所定値を
超えると、リリーフ弁３５が開いて後輪１１への過度な
トルク伝達が抑制される。9 Here, since the effective diameter difference between front and rear wheels 5 and 11 is a constant,
The rotational speed difference between the two wheels 5 and 11 increases in direct proportion to the running speed. Then, the engagement force of the clutch 23 increases according to the difference in the discharge amount (suction amount) of the first and second fluid pressure pumps 21 and 22, but the difference in the discharge amount (suction amount) increases, and the clutch When the hydraulic pressure acting on the hydraulic pressure chamber 30 exceeds a predetermined value, the relief valve 35 opens and excessive torque transmission to the rear wheel 11 is suppressed.
ところで、前述のようにして、前・後傘歯車装置６・８
のギヤ比が互いに異なる値にされ、更に第１流体圧ポン
プ２１の吐出容量が第２流体圧ポンプ２２のそれより小
さく設定されており、差動制限力が高速になるほど減少
するようになっているので、両輪５・１１の有効径差に
起因する回転速度差が増大しても、無用な伝達トルクの
増大を生じないで済む。By the way, as mentioned above, the front and rear bevel gear devices 6 and 8
The gear ratios of the first fluid pressure pump 21 are set to different values, and the discharge capacity of the first fluid pressure pump 21 is set smaller than that of the second fluid pressure pump 22, so that the differential limiting force decreases as the speed increases. Therefore, even if the rotational speed difference due to the difference in effective diameter between the two wheels 5 and 11 increases, unnecessary increases in transmitted torque can be avoided.
これらのことから、前後輪の有効径の差に起因する回転
速度差が、定速走行時における第１流体圧ポンプ２１の
吐出量が第２流体圧ポンプ２２の０
それを超えない範囲内にありさえすれば、後輪への無用
なトルク伝達を生ぜずに済む、と言うことができる。From these facts, the difference in rotational speed caused by the difference in effective diameter between the front and rear wheels is such that the discharge amount of the first fluid pressure pump 21 during constant speed driving does not exceed 0 of the second fluid pressure pump 22. It can be said that as long as there is one, there will be no need for unnecessary torque transmission to the rear wheels.
また、従駆動輪である後輪１１の有効径がより小さい場
合には、後輪１１の回転速度がより高くなるが、これは
前述した前輪５が制動されている状態と概ね等価と見做
し得るので、当然、前・後輪５・１１の連結が断たれて
トルク伝達がなされないため、前後輪の有効径差によっ
て実害を生ずる虞れは全くない。Furthermore, when the effective diameter of the rear wheel 11, which is the slave drive wheel, is smaller, the rotational speed of the rear wheel 11 becomes higher, but this is considered to be roughly equivalent to the state in which the front wheel 5 described above is braked. Naturally, the front and rear wheels 5 and 11 are disconnected and no torque is transmitted, so there is no risk of actual damage caused by the difference in effective diameter between the front and rear wheels.
以上説明したように、本発明の構成においては、第１・
第２両流体圧ポンプ２１・２２の実質的な吐出容量は、
前輪５に連動駆動される第１流体圧ポンプ２１の方がよ
り小さくなっている。そのため、クラッチの作動油圧室
３０に作用する油圧（伝達トルク）は、前・後輪５・１
１が同一径であれば、前輪５と後輪１１との回転速度が
等しくなる以前に消滅する。これは前後輪間の差動制限
力の設定が幾分か低目となっていることに相当するが、
実際には第４図に示したように、ある伝達１
トルクを発生するに要する前輪・後輪間の回転速度差は
、車速が低いほど小さくなるので、発進時など後輪へも
駆動トルクを伝達しなければならない機会が多い低車速
時には、前輪・後輪間の回転速度差に対して十分な駆動
トルクが後輪に伝達される。As explained above, in the configuration of the present invention, the first
The substantial discharge capacity of the second fluid pressure pumps 21 and 22 is as follows:
The first fluid pressure pump 21 that is driven in conjunction with the front wheels 5 is smaller. Therefore, the hydraulic pressure (transmission torque) acting on the hydraulic pressure chamber 30 of the clutch is limited to the front and rear wheels 5 and 1.
1 have the same diameter, they disappear before the rotational speeds of the front wheels 5 and rear wheels 11 become equal. This corresponds to the setting of the differential limiting force between the front and rear wheels being somewhat low,
In reality, as shown in Figure 4, the rotational speed difference between the front and rear wheels required to generate a certain transmission torque becomes smaller as the vehicle speed decreases, so drive torque is also applied to the rear wheels when starting off. At low vehicle speeds, when there are many opportunities for transmission, sufficient drive torque is transmitted to the rear wheels to compensate for the difference in rotational speed between the front and rear wheels.
なお、車速の増大に応じてクラッチ２３の伝達トルクを
減じる手段としては、上記した第１・第２両流体圧ポン
プ２１・２２のチャンバ容積を互いに異なるものとする
か、あるいは流体圧ポンプと車輪との速比を前後で異な
るものとするかをそれぞれを単独で実施しても良いし、
本実施例のように両者を組合わせても良い。Note that as a means for reducing the transmission torque of the clutch 23 in accordance with an increase in vehicle speed, the chamber volumes of the first and second fluid pressure pumps 21 and 22 may be made different from each other, or the fluid pressure pump and the wheel You can choose to have different speed ratios at the front and rear, or you can implement each separately.
The two may be combined as in this embodiment.
第５図及び第６図は、本発明の変形実施例を示しており
、上記第１図及び第２図に示した実施例と共通する部分
には同一の符号を付し、異なる部分についてのみ以下に
説明する。5 and 6 show modified embodiments of the present invention, parts common to the embodiment shown in FIGS. 1 and 2 above are given the same reference numerals, and only different parts are given. This will be explained below.
上記実施例においては、作動油圧室３０とオイルタンク
３６との間を、クラッチ作動油圧供給通路３７ｂから分
岐した通路にて連通させるものと２
しているが、本実施例においては、作動油圧室３０とオ
イルタンク３６との間に、オリフィス３８を備えた別の
連通路４７を設けるものとしている。In the above embodiment, the working hydraulic pressure chamber 30 and the oil tank 36 are communicated with each other through a passage branched from the clutch working hydraulic pressure supply passage 37b, but in this embodiment, the working hydraulic pressure chamber 30 and the oil tank 36, another communication passage 47 including an orifice 38 is provided.
これによれば、クラッチ作動油圧供給通路３７ｂから作
動油圧室３０に圧油を供給する際に、作動油圧室３０内
の空気を速やかに排出できるので、クラッチ２３の作動
応答性をより一層向上することができる。According to this, when supplying pressure oil from the clutch actuation oil pressure supply passage 37b to the actuation oil pressure chamber 30, the air in the actuation oil pressure chamber 30 can be quickly discharged, so that the actuation response of the clutch 23 is further improved. be able to.
本実施例の場合も、オリフィス３８を介してのリリーフ
流量によってクラッチ作動油圧の特性が定まること、並
び・にクラッチ２３の作動要領は、上記第１の実施例と
同様である。In the case of this embodiment as well, the characteristics of the clutch operating oil pressure are determined by the relief flow rate through the orifice 38, and the operating procedure of the clutch 23 is the same as in the first embodiment.
［発明の効果］
このように本発明によれば、前・後両輪の有効径が互い
に異なる状態で走行する際にも、従駆動輪への無用なト
ルク伝達を生ぜずに済むので、従駆動輪へ駆動トルクを
伝達する経路を構成する各部材の耐疲労強度を実質的に
低減することができる。従って、従駆動輪側部材の軽量
化を推進する上に多大な効果を奏することができる。し
かも発３
進加速時など、駆動トルクを伝達する機会が多い低速時
には、前後輪の回転速度差に応じて十分な駆動トルクが
伝達されるため、４輪駆動車輌としての実用性能を損な
わずに済む。[Effects of the Invention] As described above, according to the present invention, even when driving with the front and rear wheels having different effective diameters, unnecessary torque transmission to the driven wheels can be avoided. The fatigue strength of each member constituting the path for transmitting drive torque to the wheels can be substantially reduced. Therefore, a great effect can be achieved in promoting weight reduction of the subordinate drive wheel side member. Furthermore, at low speeds where there are many opportunities to transmit drive torque, such as during acceleration, sufficient drive torque is transmitted according to the difference in rotational speed between the front and rear wheels, so the practical performance of a four-wheel drive vehicle is not compromised. It's over.
第１図は、本発明に基づく４輪駆動車輌の動力伝達系の
全体的な構成を示すスケルトン図であり、第２図は、後
退状態にある時の油圧回路図である。
第３図及び第４図は、本発明装置の特性を示すグラフで
ある。
第５図は、本発明の変形実施例を示す第１図と同様なス
ケルトン図であり、第６図は、同変形実施例を第２図の
状態に対応して示す油圧回路図である。
１・・・エンジン、２・・・変速機、３・・・差動装置
、４・・・ドライブシャフト、５・・・前輪、６・・・
傘歯車装置、７・・・動力伝達装置、８・・・傘歯車装
置、９・・・差動装置、１０・・・ドライブシャフト、
１１・・・後輪、２１・・・第１流体圧ポンプ、２２・
・・第２流体圧ポンプ、２３・・・流体圧作動クラッチ
（トルク伝達装置）、４
２４・・・第１ポート、２５・・・第２ポート、２６・
・・第３ボート、２７・・・第４ポート、２８・・・第
１連結油路、２９・・・第２連結油路、３０・・・作動
油圧室、３１・・・切換弁、３２・・・第１弁室、３３
・・・第２弁室、３４・・・一方向弁、３５・・・リリ
ーフ弁（伝達トルク制限手段）、３６・・・オイルタン
ク、３７ａ・・・バイパス油路、３７ｂ・・・作動油圧
供給油路、３８・・・オリフィス、４７・・・連通路
特 許 出 願 人 本田技研工業株式会社代 理
人 弁理士 大 島 陽 −（外−名）
５
４聰上λへFIG. 1 is a skeleton diagram showing the overall configuration of a power transmission system of a four-wheel drive vehicle according to the present invention, and FIG. 2 is a hydraulic circuit diagram when the vehicle is in a backward state. 3 and 4 are graphs showing the characteristics of the device of the present invention. FIG. 5 is a skeleton diagram similar to FIG. 1 showing a modified embodiment of the present invention, and FIG. 6 is a hydraulic circuit diagram showing the modified embodiment corresponding to the state shown in FIG. 2. DESCRIPTION OF SYMBOLS 1... Engine, 2... Transmission, 3... Differential device, 4... Drive shaft, 5... Front wheel, 6...
Bevel gear device, 7... Power transmission device, 8... Bevel gear device, 9... Differential device, 10... Drive shaft,
DESCRIPTION OF SYMBOLS 11... Rear wheel, 21... First fluid pressure pump, 22.
...Second fluid pressure pump, 23...Fluid pressure operated clutch (torque transmission device), 4 24...First port, 25...Second port, 26...
... Third boat, 27... Fourth port, 28... First connecting oil path, 29... Second connecting oil path, 30... Working hydraulic chamber, 31... Switching valve, 32 ...First valve chamber, 33
...Second valve chamber, 34...One-way valve, 35...Relief valve (transmission torque limiting means), 36...Oil tank, 37a...Bypass oil passage, 37b...Working oil pressure Supply oil passage, 38... Orifice, 47... Communication path Patent Applicant: Honda Motor Co., Ltd. Agent Patent attorney: Akira Oshima - (foreign name) 5 4 To λ
Claims (1)
する第２部材と、前記第１部材と前記第２部材との間に
介設された前記前輪と前記後輪との回転速度差に応じて
伝達トルクが変化するトルク伝達装置とを有する４輪駆
動車輌の動力伝達装置であって、 前記トルク伝達装置の伝達トルクの上限を規定する伝達
トルク制限手段と、 前記トルク伝達装置の伝達トルクを車速の増大に応じて
減じる手段とを有することを特徴とする４輪駆動車輌の
動力伝達装置。(1) A first member that rotates in conjunction with the front wheel, a second member that rotates in conjunction with the rear wheel, and rotation of the front wheel and the rear wheel that are interposed between the first member and the second member. A power transmission device for a four-wheel drive vehicle, comprising: a torque transmission device whose transmission torque changes according to a speed difference; a transmission torque limiting means for defining an upper limit of the transmission torque of the torque transmission device; and the torque transmission device. 1. A power transmission device for a four-wheel drive vehicle, comprising: means for reducing the transmitted torque according to an increase in vehicle speed.
Priority Applications (3)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| GB9024529A GB2239921B (en) | 1989-11-15 | 1990-11-12 | Power transmission apparatus for a four-wheel drive vehicle |
| US07/612,766 US5219038A (en) | 1989-11-15 | 1990-11-14 | Power transmission device for a four-wheel drive vehicle |
| DE4036280A DE4036280C2 (en) | 1989-11-15 | 1990-11-14 | Power transmission system for a four-wheel drive vehicle |
Applications Claiming Priority (2)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP1-296639 | 1989-11-15 | ||
| JP29663989 | 1989-11-15 |
Publications (2)
| Publication Number | Publication Date |
|---|---|
| JPH03224831A true JPH03224831A (en) | 1991-10-03 |
| JP2963174B2 JP2963174B2 (en) | 1999-10-12 |
Family
ID=17836142
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| JP23824590A Expired - Lifetime JP2963174B2 (en) | 1989-11-15 | 1990-09-07 | Power transmission device for four-wheel drive vehicle |
Country Status (1)
| Country | Link |
|---|---|
| JP (1) | JP2963174B2 (en) |
Cited By (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| EP0850798A1 (en) | 1995-10-18 | 1998-07-01 | Honda Giken Kogyo Kabushiki Kaisha | Slave driving force-transmitting mechanism for vehicle |
| US5951401A (en) * | 1996-12-31 | 1999-09-14 | Honda Giken Kogyo Kabushiki Kaisha | Slave driving force-transmitting mechanism for vehicle |
| US6694592B2 (en) | 1999-09-30 | 2004-02-24 | Honda Giken Kogyo Kabushiki Kaisha | Method for tightening fixing band and constant velocity universal joint apparatus |
-
1990
- 1990-09-07 JP JP23824590A patent/JP2963174B2/en not_active Expired - Lifetime
Cited By (5)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| EP0850798A1 (en) | 1995-10-18 | 1998-07-01 | Honda Giken Kogyo Kabushiki Kaisha | Slave driving force-transmitting mechanism for vehicle |
| US5951401A (en) * | 1996-12-31 | 1999-09-14 | Honda Giken Kogyo Kabushiki Kaisha | Slave driving force-transmitting mechanism for vehicle |
| US6694592B2 (en) | 1999-09-30 | 2004-02-24 | Honda Giken Kogyo Kabushiki Kaisha | Method for tightening fixing band and constant velocity universal joint apparatus |
| US6725530B2 (en) | 1999-09-30 | 2004-04-27 | Honda Giken Kogyo Kabushiki Kaisha | Apparatus and method for tightening fixing band and constant velocity universal joint apparatus |
| US7065862B1 (en) | 1999-09-30 | 2006-06-27 | Honda Giken Kogyo Kabushiki Kaisha | Apparatus for tightening fixing band and constant velocity universal joint apparatus |
Also Published As
| Publication number | Publication date |
|---|---|
| JP2963174B2 (en) | 1999-10-12 |
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