fig1 shows a schematic view of an attachment 1 in a vehicle ( not shown in detail ), for example a toolbar of a tractor . the attachment 1 has a hydraulic cylinder 2 , which in the present embodiment is made as a single - acting cylinder . however , it is also possible to use a double - acting cylinder instead of a single - acting cylinder 2 . a control device 3 , which will be explained in detail in the following , controls the cylinder 2 . a load 4 is attached to attachment 1 to show that the attachment must be able to lift a corresponding load . the load 4 can , for example , be a plough , which must be pulled by a tractor . this plough must have a certain penetration depth into the earth . this penetration depth is achieved in that the attachment 1 is set to a certain height position , in the following called “ position ”. in fig1 , this position is shown by “ x ”. however , it is not sufficient to set this position once . when the tractor changes its driving direction , and the plough has to throw the blocks in different directions , or when the tractor leaves the field , the plough must be lifted . after reassuming the work , the plough has to be lowered to the desired position again . for this purpose , the control device 3 has a 3 - way , 3 - direction control valve 5 in the form of a proportional valve . by means of the control valve 5 it is possible to let hydraulic fluid flow from a pressure connection p to the inlet a of the hydraulic cylinder 2 to lift the load . or the load 4 can be lowered by means of the control valve 5 by connecting the inlet a of the cylinder 2 with a tank outlet r 1 . alternatively , the control valve 5 can also be divided into two valves , as known , for example , from u . s . pat . no . 6 , 058 , 343 . in this case , one valve serves the purpose of lifting the load 4 and one serves the purpose of lowering the load 4 . in the case of a double - acting cylinder , a valve can be used , which controls four ways and three directions . also such a valve can be divided into two valve parts or four valve parts for the separate control of the individual functions , as known from u . s . pat . no . 5 , 960 , 695 . the hydraulic system , which comprises the cylinder 2 and the control device 3 , has a pilot - controlled non - return valve 6 , which can be used , when it is desired to avoid a leakage from the connection a , so - called zero - leakage . in this connection , fig1 shows that a pilot pressure pf is supplied to the pilot - controlled non - return valve 6 via the control valve 5 . alternatively , the pilot pressure can be supplied directly or via a solenoid valve . this causes that the non - return valve 6 can be controlled independently of the position of the control valve 5 . this may be advantageous in cases involving a control , in which the dynamics of the pilot - controlled non - return valve 6 plays a role . for the inlet control is used a pressure control valve 7 , which produces a constant pressure drop over the control valve 5 . thus , in wide limits , a load - independent lifting behaviour of the cylinder 2 can be achieved . for the outlet control is used a pressure sensor 8 and a microcontroller 9 . the pressure sensor 8 produces an output signal , which can be evaluated by the microcontroller 9 . the pressure compensation can be made in different ways , for example by means of a feedback linearisation or with a flow estimation function and control . the microcontroller 9 also serves as position sensor , that is , it detects the position “ x ”. this is shown schematically by a line from “ x ” to the microcontroller 9 . fig2 shows the principle of the feedback linearisation . the measured pressure drop δp , that is , the pressure difference between the connection a of the cylinder 2 and the outlet r 1 , is led back to an inverted model 10 of the valve 5 , which results in a linearised feedback system , in which the desired flow q r is equal to the flow q , independently of the load pressure . k is a valve constant . fig3 shows a model , which uses an estimation function to control the outlet of fluid from the cylinder 2 . the estimated flow q est is fed back and compared with the specified reference flow q r in a controller r . this flow is simply estimated in that the pressure difference δp between the connection a of the cylinder 2 and the return connection r 1 , and the valve control signal u , are converted with a throttling behaviour . also in this case it can be achieved with good proximity that the flow q leaving the cylinder 2 is in fact equal to the desired flow q r . when a double - acting cylinder is used instead of the hydraulic cylinder 2 , and for this reason the inlet control and the outlet control is separated , for example with two or four control valves , then the outlet pressure and the inlet pressure must be controlled to avoid that the load 4 runs away . however , in many cases the separation is not required . when a 4 - way , 3 - direction valve ( 4 / 3 - valve ), which is controlled on the basis of the flow ( meter - in flow controlled valve ), is used to control the double - acting cylinder , the flow in both directions is proportional to the control signal , provided that the valve has been dimensioned correctly for the maximum load . in this case a return - flow measuring ( meter - out ) is not required . as the flow control device produces a linear function between the desired flow value inlet q r and the speed of the load dx / dt , the fastest movement time , as shown in fig4 , between the positions x 0 and xf is the time t min ( in seconds ), v max being the maximum speed at completely open valve 5 and a load pressure δp . when a lower speed v set is desired , the opening time t set of the valve 5 must be longer to achieve the same movement . theoretically , the linear profiles shown require an indefinitely large acceleration . in practice , the pressure δp limits the acceleration , with large loads it may , however , happen anyway that the front wheels leave the ground , because the inertia forces are too large . for this reason , an acceleration limit is introduced , as shown in fig4 . the acceleration profile has a maximum a and a minimum − a . the position of the load is simply found in that the acceleration profile is integrated twice and the desired speed v set is included , that is , is used as integration limit . this gives a controlled acceleration , which , however , causes a somewhat longer movement time tf . the profile is optimal in that it gives the shortest movement time for a predetermined maximum acceleration and a maximum speed . the speed can also be replaced by a time specification . a further integration will convert the speed profile to a trajectory for the movement of the load 4 or the attachment 1 , respectively . this is obvious to a person skilled in the art and therefore not shown in detail . when other acceleration limits are chosen , the speed course will change . however , in any case it can be avoided that a maximum acceleration is exceeded . when a monitoring of both acceleration and speed is wanted , the desired trajectory gets somewhat more complicated . for this reason , a follower is expedient . fig5 shows a first embodiment of such a follower . as described above , a trajectory generator 11 produces a trajectory , that is , the individual positions x over the time , as shown , for example in fig4 . the control has a directly connected part , which represents an inverted model 12 of the attachment 1 . the transfer function from x r ( specification ) to x will thus be 1 . this does not necessarily mean that x r = x , and also not that x approaches the course x r . as , however , the attachment system itself is unstable , a position control with a controller c is additionally used . the immediate position x of the load 4 is supplied to this controller c . the effect of this measure is that x approaches the specification x r and that the control system gets stabler in relation to parameter variations and interferences . in a preferred embodiment , the model approaches the attachment with a constant and an integrator in such a manner that the inverted model becomes a constant and a differentiator . the approach simplifies the control and is sufficiently accurate for moderate accelerations . the controller c is dimensioned so that the control system has a predetermined stability area . it is expedient to arrange a dead - band compensator 14 between the outlet control 13 , as shown , for example , in fig2 or fig3 , and the system of the attachment 1 , in order to compensate a dead - band db , which is required in the valve 5 to let the load - sensing and pilot pressure signals pass . of course , the system described can be used in both directions , that is , both when lifting and when lowering a load . during lifting , the outlet control 13 will play no role . an alternative method for combining an outlet control and a follower is shown in fig6 . this embodiment comprises an adaptive follower . in this case an outlet control is not absolutely necessary , when the load is constant or changes slowly . the adaptive follower adapts the inverted model of the system of the attachment 1 .
the invention concerns a vehicle with an attachment , which has a lifting drive with at least one hydraulic cylinder and a control device , the control device having a position sensor and a position generator . it is endeavored to improve the position control of the attachment . for this purpose , the control device has a trajectory generator , which generates a trajectory in dependence of a desired position value and an acceleration limit , said trajectory being optimized with regard to a time specification , and a follower , which controls the lift drive in dependence of the trajectory .