METHOD AND SYSTEM FOR A FLOW-ISOLATED VALVE ARRANGEMENT AND A THREE-CHAMBER CYLINDER HYDRAULIC ARCHITECTURE

A hydraulic circuit is disclosed which includes one or more i) linear; or ii) rotary hydraulic actuator, wherein total number of cylinder chambers is N, M pressure rails, a valve arrangement, including M hydraulic rail ports each coupled to a pressure rail, N hydraulic chamber ports each coupled to a chamber of one or more actuators, N proportional valves each corresponding to one of the N hydraulic chamber ports, X sets of on-off valves and check valves coupling two or more hydraulic rail ports to each of supply sides of each of the N proportional valves, and Y sets of on-off valves and check valves coupling two or more hydraulic rail ports to each of return sides of each of the N proportional valves, and a controller configured to in real-time operate the N proportional valves and the associated on-off valves to achieve one or more desired functional parameters.

STATEMENT REGARDING GOVERNMENT FUNDING

TECHNICAL FIELD

The present disclosure generally relates to hydraulic architectures, and in particular, to a three-chamber cylinder hydraulic architecture specifically useful in construction machinery as well as a flow-isolated valve arrangement.

BACKGROUND

Hydraulic systems utilized in heavy machinery are quite well known. In early days, a simple hydraulic cylinder was utilized to generate a force to move objects based on the hydraulic pressure within the cylinder and the effective area of a piston moving within the cylinder resulting in a load force. In typical cylinders, two chambers are used, each with a respective effective area, such an arrangement creates a force in each direction. The net force resultant of pressurized fluid acting in both areas of the cylinders is typically referred to as a cylinder load. Traditional applications usually act on the flowrate in/out one chamber while the remainder chamber is kept at a pressure as low as possible in order to reduce system losses.

Although the initial concept was quite simple, different hydraulic control architectures were developed over the years, especially as the number of hydraulic actuators per machine increased. Commonly, these architectures have the use of shared hydraulic power supply, e.g., a hydrostatic pump, and control valves dedicated to each actuator. The associated challenges with these architectures are twofold: first, controllability of multiple actuators with a single and shared source of hydraulic power; and second, energy efficiency. While different approaches in the prior art have been successful with regards to the first challenge, most circuits currently available in the market still suffer from low efficiency when powering more than one actuator at a time.

When more than one hydraulic actuator shares the same hydraulic supply, the supplied pressure must be slightly higher than the maximum pressure requirements in the system. Therefore, any other hydraulic actuator requiring lower pressure to achieve the desired load will need throttle control to decrease the supply pressure to the desired level, which leads to power losses. In the present disclosure the term “hydraulic actuator” refers to either rotary or linear actuators. Consequently, any system aiming at energy efficiency will need to minimize the pressure difference between the supply system and the pressure requirements of each one of the multiple actuators sharing the same supply.

To achieve such a condition, it is possible to 1) increase the number of supply pressure rails such that more than one supply pressure is available and 2) increase the number of cylinder chambers such that different combinations of connections between chambers and supply rails can be used to minimize the throttling requirement, therefore reducing system losses. In short, with more options to combine different chamber areas and pressures, it is possible to achieve the same load (effective cylinder force) with a lower difference between supply pressure and chamber pressure, therefore decreasing throttling losses. As a consequence, the higher the number of available combinations, (cylinder modes) the more efficient the cylinder should be, in theory.

The relationship between pressure rails and the number of chambers in achieving discretized number of possible connections between pressure rails and cylinder chambers (modes) for different applications is governed by:

Number of Discrete Modes=(Number of Chambers)Number of pressure Rails. These different combinations of connections between supply pressures and cylinder chambers are also sometimes referred to as discrete force levels available to the actuator.

An example of such a multi-chamber arrangement is provided in U.S. Pat. No. 10,704,569 to Sipola et al., in which at least a four-chamber actuator was introduced utilizing two pressure rails identified as high pressure (HP) and low pressure (LP). As shown inFIG.1a, each chamber is coupled to each of the pressure rails utilizing a system of valves in parallel. That is, each chamber is coupled to both the high-pressure rail and the low-pressure rail utilizing two separate valves, one between each pressure rail and the chamber. In this case, proportional valves were utilized to synchronize valves opening and closing, but ultimately, non-throttle control was used. Therefore, the proportional valves are kept either fully opened or fully closed, (except during transitions) since any partial opening of the valves would introduce fluid throttling, characterizing throttle control.

Based on the formula provided above, in the example shown inFIG.1a, the number of discretized forces is 42which equals 16, as shown inFIG.1b, which is a graph of discretized force per an index spanning from 1 to 16. This means that, if non-throttled control is to be used, as described in the '569 patent, only 16 load forces are achievable, limiting the accuracy of speed tracking, specially at low speeds. This is because any mismatch between the load force and the actual actuator force requirement will result in a cylinder acceleration. Therefore, if precise motion control is to be achieved, continuous throttling in at least one of the chambers is required, therefore creating a trade-off between control accuracy and efficiency.

The approach shown in the '569 patent is typical in the prior art (see, e.g., WO 2014081353 A1). However, there are disadvantages with these approaches. For example, the number of chambers in a cylinder and the number of pressure rails (two in the '569 patent) result in a reduced number of discretized forces, while requiring complicated cylinder designs given the higher number of chambers. As mentioned above, in the '569 patent the combination of 4 chambers and two pressure rails resulted in 16 discretized forces. In the WO 2014081353 A1 publication, a five-chamber actuator was used with two pressure rails which can result in 25 discretized forces. However, in all these iterations only two pressure rails were used resulting in a costly and complicated cylinder arrangement with a reduced number of discretized forces. Including an extra pressure rail in the mentioned architectures would also not be cost-effective since a large number of valves would be required.

In addition, both references mentioned use of non-throttle control, or a mix of on/off and proportional valves, which limits achievable efficiency and performance of such systems.

Another aspect to be considered, is that in these types of architectures the cylinder controller constantly changes the supply line connected to each chamber. This means that during a short transient period one valve (i.e., connecting the chamber to the high-pressure line) will be closing while the other one (i.e., connecting the same chamber to the low-pressure line) will be opening. Since these valves are not infinitely fast, both valves will be opened for a short period of time, creating a short circuit between high and low-pressure supply lines. This ultimately causes significant leakages and lowers the system efficiency.

The described short-circuit phenomenon exists regardless of the use of on/off or proportional valves. The rationale for using proportional valves in the mentioned prior art is based on the need to synchronize and delay opening and closing of the valves connecting different rails to a given chamber. Such approach is able to reduce the effect of this short-circuit in the system efficiency. Nevertheless, such a solution is not able to eliminate the problem. In addition, such a solution requires proportional valves with very short closing times and since throttle control was not used the cylinder is still limited to a finite number of available forces.

Therefore, there is an unmet need for a novel approach in hydraulic architecture and its control method such that a more precise motion control can be achieved without significant increase in system losses due to throttle control as well as without increase in cylinder design complexity as well as a novel method and system approach that can provide an isolated flow from pressure rails without a short circuit between any two pressure rails when one pressure rail is switched to the other pressure rail.

SUMMARY

According to one embodiment a valve arrangement is disclosed which includes M hydraulic rail ports each configured to be coupled to a pressure rail, N hydraulic chamber ports each configured to be coupled to a chamber of one or more actuators. N proportional valves each corresponding to one of the N hydraulic chamber ports. Each proportional valve includes a rail side coupled to the M hydraulic rail ports and a chamber side coupled to a corresponding hydraulic chamber port. Each rail side of the N proportional valves is divided into a supply side configured to supply hydraulic fluid to a corresponding hydraulic chamber port and a return side configured to receive hydraulic fluid from the corresponding hydraulic chamber port. The valve arrangement further includes X sets of on-off valves and check valves coupling two or more hydraulic rail ports to each of the supply sides of each of the N proportional valves, and Y sets of on-off valves and check valves coupling two or more hydraulic rail ports to each of the return sides of each of the N proportional valves. Selectively operating each of the on-off valves and the proportional valves provides selective pressure or flow to each one of the N hydraulic chamber ports.

According to one embodiment, in the above valve arrangement X has a maximum number of M−1.

According to one embodiment, in the above valve arrangement X has a minimum number of 1.

According to one embodiment, in the above valve arrangement Y has a maximum number of M−1.

According to one embodiment, in the above valve arrangement Y has a minimum number of 1.

According to one embodiment, in the above valve arrangement the on-off valves and the check valves on the supply side of each of the N proportional valves cooperate to selectively define a pressure in the supply side of the proportional valve.

According to one embodiment, in the above valve arrangement the on-off valves and the check valves on the supply side of each of the N proportional valves cooperate to prevent fluid flow between a hydraulic rail port with a first pressure to a hydraulic rail port with a second pressure, wherein the first pressure is higher than the second pressure.

According to one embodiment, in the above valve arrangement the on-off valves and the check valves on the return side of each of the N proportional valves cooperate to selectively define a pressure in the return side of the proportional valve.

According to one embodiment, in the above valve arrangement the on-off valves and the check valves on the return side of each of the N proportional valves cooperate to prevent fluid flow between a hydraulic rail port with a first pressure to a hydraulic rail port with a second pressure, wherein the first pressure is higher than the second pressure.

According to another embodiment, a hydraulic circuit is also disclosed which includes one or more i) linear; or ii) rotary hydraulic actuator each with one or more cylinder chambers disposed therein, wherein total number of cylinder chambers is N, M pressure rails, each at a corresponding pressure, and a valve arrangement. The valve arrangement includes M hydraulic rail ports each configured to be coupled to a pressure rail, N hydraulic chamber ports each configured to be coupled to a chamber of one or more actuators, N proportional valves each corresponding to one of the N hydraulic chamber ports, wherein each proportional valve includes a rail side coupled to the M hydraulic rail ports and a chamber side coupled to a corresponding hydraulic chamber port, and wherein each rail side of the N proportional valves is divided into a supply side configured to supply hydraulic fluid to a corresponding hydraulic chamber port and a return side configured to receive hydraulic fluid from the corresponding hydraulic chamber port, X sets of on-off valves and check valves coupling two or more hydraulic rail ports to each of the supply sides of each of the N proportional valves, and Y sets of on-off valves and check valves coupling two or more hydraulic rail ports to each of the return sides of each of the N proportional valves. Selectively operating each of the on-off valves and the proportional valves provides selective pressure or flow to each one of the N hydraulic chamber ports. The hydraulic circuit also includes a controller configured to receive one or more desired functional parameters for the one or more cylinder chambers and in real-time i) receive data from a plurality of sensors associated with the one or more cylinder chambers, and ii) activate and deactivate the N proportional valves and the associated on-off valves to achieve the one or more desired functional parameters.

According to one embodiment, in the above hydraulic circuit X has a maximum number of M−1.

According to one embodiment, in the above hydraulic circuit X has a minimum number of 1.

According to one embodiment, in the above hydraulic circuit Y has a maximum number of M−1.

According to one embodiment, in the above hydraulic circuit Y has a minimum number of 1.

According to one embodiment, in the above hydraulic circuit the on-off valves and the check valves on the supply side of each of the N proportional valves cooperate to selectively define a pressure in the supply side of the proportional valve.

According to one embodiment, in the above hydraulic circuit the on-off valves and the check valves on the supply side of each of the N proportional valves cooperate to prevent fluid flow between a hydraulic rail port with a first pressure to a hydraulic rail port with a second pressure, wherein the first pressure is higher than the second pressure.

According to one embodiment, in the above hydraulic circuit the on-off valves and the check valves on the return side of each of the N proportional valves cooperate to selectively define a pressure in the return side of the proportional valve.

According to one embodiment, in the above hydraulic circuit the on-off valves and the check valves on the return side of each of the N proportional valves cooperate to prevent fluid flow between a hydraulic rail port with a first pressure to a hydraulic rail port with a second pressure, wherein the first pressure is higher than the second pressure.

According to one embodiment, in the above hydraulic circuit each of the M pressure rails is sourced from one or more power sources.

According to one embodiment, in the above hydraulic circuit the power source is an internal combustion engine.

According to one embodiment, in the above hydraulic circuit the power source is one or more electric motors.

According to one embodiment, in the above hydraulic circuit the pressures in the pressure rails are kept at the desired levels by one or more hydrostatic pumps of either fixed or variable displacement.

According to one embodiment, in the above hydraulic circuit real-time measured states including pressure, force, torque, position and speed are used to adjust desired pressure levels and associated variation range in the pressure rails.

According to one embodiment, in the above hydraulic circuit the one or more functional parameters includes force.

According to one embodiment, in the above hydraulic circuit the one or more functional parameters includes speed.

According to one embodiment, in the above hydraulic circuit the one or more functional parameters includes position.

According to one embodiment, in the above hydraulic circuit the controller controls the N proportional valves and the associated on-off valves based on minimizing energy losses between the supply side and the return side of each of the N proportional valves.

According to one embodiment, in the above hydraulic circuit the controller utilizes the data from the plurality of sensors associated with the one or more cylinder chambers in one or more feedback loops.

A hydraulic force generator for use with heavy machinery is also disclosed which consists of a hydraulic actuator with three chambers disposed therein; three hydraulic pressure rails consisting of i) a high-pressure rail, ii) a medium pressure rail, and iii) a low pressure rail; and at least 3·N−M proportionally controlled hydraulic valves coupled to the hydraulic linear actuator, wherein each chamber is coupled to N hydraulic pressure rails via proportional valves, wherein continuous force control is achieved by proportionally controlling the opening area of each valve. M is the number of optionally removable valves and 0≤M≤2N-2.

According to one embodiment, in the above hydraulic force generator the N hydraulic pressure rails are sourced from a single power source.

According to one embodiment, in the above hydraulic force generator the single power source is an internal combustion engine.

According to one embodiment, in the above hydraulic force generator the single power source is one or two electric motors powered by a battery pack.

According to one embodiment, in the above hydraulic force generator each of the N hydraulic pressure rails represents hydraulic power supplied by a single hydrostatic pump, having an outlet serving each of the N hydraulic pressure rails through a directional valve.

According to one embodiment, in the above hydraulic force generator two or more hydrostatic pumps are used to supply hydraulic power to the N hydraulic pressure rails.

According to one embodiment, in the above hydraulic force generator the hydrostatic pump(s) is based on one of fixed or variable displacement.

According to one embodiment, in the above hydraulic force generator N is 3.

According to one embodiment, in the above hydraulic force generator N is 2.

According to one embodiment, in the above hydraulic force generator real-time measured states including pressure, force, position and speed are used to adjust desired pressure levels and associated variation range in the pressure rails.

A hydraulic control system for use with heavy machinery is also disclosed which includes a hydraulic actuator with three chambers disposed therein, three hydraulic pressure rails consisting of i) a high-pressure rail, ii) a medium pressure rail, and iii) a low-pressure rail, and at least 3·N−M proportionally controlled hydraulic valves coupled to the hydraulic actuator, wherein each chamber is coupled to N hydraulic pressure rails via proportional valves, wherein continuous force control is achieved by proportionally controlling the opening area of each valve. M is the number of optionally removable valves and 0≤M≤2N-2. The hydraulic control system also includes a control unit responsible for adjusting the proportional valves opening such that pressure closed loop-pressure control by means of fluid throttling can be achieved in each one of the multi-chamber cylinder chambers. Such pressure controller can also be used as the inner-loop of a closed-loop speed or position control.

According to one embodiment, in the above hydraulic control system the multi-chamber cylinder includes pressure sensors in the hydraulic lines up and downstream each proportional valve.

According to one embodiment, in the above hydraulic control system position or speed sensors are included such that closed-loop position/speed control can be achieved.

According to one embodiment, in the above hydraulic control system the N hydraulic pressure rails are sourced from a single power source.

According to one embodiment, in the above hydraulic control system the power source is an internal combustion engine.

According to one embodiment, in the above hydraulic control system the power source is one or two electric motors powered by a single battery pack.

According to one embodiment, in the above hydraulic control system each of the N hydraulic pressure rails represents hydraulic power supplied by a single hydrostatic pump, having an outlet serving each of the N hydraulic pressure rails through a directional valve.

According to one embodiment, in the above hydraulic control system two or more hydrostatic pumps are used to supply hydraulic power to the N hydraulic pressure rails.

According to one embodiment, in the above hydraulic control system the hydrostatic pump(s) is based on one of fixed or variable displacement.

According to one embodiment, in the above hydraulic control system N is 3.

According to one embodiment, in the above hydraulic control system N is 2.

DETAILED DESCRIPTION

A novel valve arrangement in hydraulic architectures is provided herein that can provide independent chamber pressure control with an isolated flow from pressure rails without a short circuit between any two pressure rails when one pressure rail is switched to the other pressure rail. Additionally, a novel method and system approach in hydraulic architectures is provided herein that utilizes the aforementioned novel valve arrangement. Towards this end, reference is made toFIG.2in which a schematic of valve arrangement100according to the present disclosure is shown. In the embodiment shown inFIG.2and all other embodiments of the present disclosure, the valve arrangement includes M hydraulic rail ports1011,1012,1013each configured to be coupled to a pressure rail (inFIG.2, M=3 including high pressure (HP)102, medium pressure (MP)104, and low pressure (LP)106), N hydraulic chamber ports each configured to be coupled to a chamber of an actuator, and N or 2N proportional valves110(inFIG.2, N=1). Each proportional valve includes a rail side112that can be coupled to up to M hydraulic rail ports and a chamber side that is coupled to a corresponding hydraulic chamber port108. When N proportional valves are used, each rail side of the N proportional valves is divided into a supply side that is configured to supply hydraulic fluid to a corresponding hydraulic chamber port and a return side configured to receive hydraulic fluid from the corresponding hydraulic chamber port. It is also possible to use 2N 2-way valves, one connecting the chamber to the supply side and the other connecting the chamber to the return side. In addition, the valve arrangements of the present disclosure include X sets of on-off valves1201,1202and check valves1221,1222that couple two or more hydraulic rail ports1011,1012,1013to each of the supply sides116of each of the N proportional valves110(inFIG.2, X=2). X has a maximum number of M−1 (in this case M=3 and X=2). X has a minimum number of 1. Furthermore, the valve arrangements of the present disclosure include Y sets of on-off valves1241,1242and check valves1261,1262that couple two or more hydraulic rail ports1011,1012,1013to each of the return sides118of each of the N proportional valves110(inFIG.2, Y=2), has a maximum number of M−1. Y has a minimum number of 1. As discussed below, selective operation of each of the on-off valves and the proportional valves provides accurate and efficient pressure-control to each one of the N hydraulic chamber ports.

Additionally, the on-off valves1201,1202and the check valves1221,1222on the supply side116of each of the N proportional valves110cooperate to selectively define a pressure in the supply side116of the proportional valve110. Furthermore, the on-off valves1201,1202and the check valves1221,1222on the supply side116of each of the N proportional valves110cooperate to prevent fluid flow between a hydraulic rail port1011,1012,1013with a first pressure to a hydraulic rail port1011,1012,1013with a second pressure, wherein the first pressure is higher than the second pressure. Yet additionally, the on-off valves1241,1242and the check valves1261,1262on the return side118of each of the N proportional valves110cooperate to selectively define a pressure in the return side118of the proportional valve110. Yet furthermore, the on-off valves1241,1242and the check valves1261,1262on the return side118of each of the N proportional valves110cooperate to prevent fluid flow between a hydraulic rail port1011,1012,1013with a first pressure to a hydraulic rail port1011,1012,1013with a second pressure, wherein the first pressure is higher than the second pressure.

As discussed above, the valve arrangement100shown inFIG.2is coupled to three pressure rails: high-pressure rail102, medium-pressure rail104, and low-pressure rail106. These three pressure rails are coupled to a chamber of an actuator via a combination of on/off valves1201,1202,1241,1242(shown as 1V2, 1V3, 1V7, and 1V9) and check valves1221,1222,1221,1222(shown as 1V4, 1V5, 1V6, and 1V8) as well as the proportional valve110(shown as 1V1). In the embodiment shown, an inlet port130of the 3/3 proportional valve110(1V1) is coupled to the high-pressure rail102through the on/off valve1201(1V2). The same port is also coupled to the medium pressure rail104through on/off valve1202(1V3) and check valve1221(1V4). The inlet port130is also coupled to the low-pressure rail106through the check valve1221(1V5). Similarly, an outlet port132of the proportional valve110(1V1) is coupled to the high-pressure rail102through the check valve1261(1V6), to the medium-pressure rail104through the on/off valve1242(1V9) and check valve1262(1V8) and to the low-pressure rail106through the on/off valve1241(1V7). A pre-loaded check valve128identified as 1V10 can be added to avoid cavitation in the chamber. Additionally, a relief valve130identified as 1V11 is used as a safety device that limits the maximum pressure in the chamber.

To better elucidate the operation of the valve arrangement100shown inFIG.2, the following scenario is provided as an example. Suppose the on/off valve1202(1V2) is turned on to initially provide high pressure from the high pressure rail102to the proportional valve110(1V1). If the supply side116desires to change the pressure to medium pressure by coupling the proportional valve110(1V1) to the medium pressure rail104, the on/off valve1201(1V2) is deactivated while at the same time the on/off valve1202(1V3) is activated. The issue of short-circuit described above with respect to the prior art is alleviated by the check valves1221(1V4) and1222(1V5) whereby high pressure fluid from the proportional valve110(1V1) is cut off from the medium pressure rail104(MP Line) and the low pressure rail106(LP Line) by way of the check valves1221(1V4) and1222(1V5), respectively.

It should be appreciated that the arrangement100shown inFIG.2is for example only. The valve arrangement of the present disclosure can include more or less number of pressure rails. It should also be noted that different valve assemblies may be possible while using the same concept. For example, depending on the application it may not be necessary to couple both inlet/outlet ports130/132of the proportional valve110(1V1) to all three pressure rails. In that scenario, the number of on/off valves may be reduced. Examples of such circuits are shown inFIGS.3a,3b, and3c, each of which provide an example of a hydraulic circuit.

FIG.3ais similar toFIG.2in that a proportional valve210(2V1) is coupled to two on/off valves220(2V2) and224(2V4).FIG.3bis similar toFIG.2in that a proportional valve310(3V1) is coupled to three on/off valves320(3V2),324(3V4), and326(3V7).FIG.3cis similar toFIG.2in that a proportional valve410(4V1) is coupled to three on/off valves420(4V2),424(4V4), and426(4V6).

Similarly, when two chambers with flow in opposite directions (i.e. a cylinder with2opposing chambers), it is possible that the two associated proportional valves coupled to each chamber share the same set of on/off and check valves, as shown inFIG.4which provides a schematic of an embodiment of the valve arrangement according to the present disclosure with2chambers with flow in opposite directions.

In the present disclosure the valve arrangement100ofFIG.2, or its possible variations within the skill set of a person having ordinary skill in the art, is used for controlling pressure within linear actuator chambers. It should be highlighted that it is also possible to expand the concept to architectures with more than 2 chambers, by replicating the configuration shown inFIG.2and/orFIG.4as necessary depending on the application needs. An example of a system with multiple chambers is presented inFIG.5, which provides a schematic of an embodiment of the valve arrangement according to the present disclosure in a system with3chambers.

With reference toFIG.5, suppose the multi-chamber cylinder is extending, with chambers A and C expanding and chamber B retracting. A control mechanism responsible for operating the valves is commanded to move proportional valve 6V10 to stay between the center and left-most position such that it connects the chamber side to the supply side of the proportional valve and the respective chamber pressure can be controlled as desired with flow from the rails to the chamber. Similarly, the proportional valve 6V15 is kept between the center and right-most positions, connecting its respective chamber side to the return side and controlling the pressure in its respective chamber, with flow from the chamber to the rails.

To minimize throttling losses across valve 6V10, a supervisory controller selects between the available pressure levels in the supply side of the proportional valve and commands the state of the on/off valves 6V11 and 6V12. Similarly, to minimize the throttling losses across the valve 6V15, the controller selects between the available pressure levels in the return side, determining the state of the on/off valves 6V17 and 6V19. The set of valves connected to chamber C are controlled in a similar fashion to those of chamber A and the remaining on/off valves 6V7 and 6V9 remain closed.

During cylinder retraction, the operation is similar. However, in this case the proportional valve 6V10 will be between the center and right-most position, connecting chamber A to the return side, while the proportional valve 6V15 will be between center and left-most position, connecting chamber B to supply side. The pressure in chamber C is controlled in a similar fashion to those of chamber A, with its own dedicated set of valves. Should a fourth chamber be added to the cylinder in the opposing direction to that of chamber B, it could also share the set of on/off and check-valves used to supply the proportional valve 6V1.

The valve arrangement of the present disclosure for controlling pressure within a multi-chamber cylinder provides several advantages over the arrangements of the prior art discussed above. First, the valve arrangement of the present disclosure avoids any short-circuit between the pressure rails when the valves are switched from one pressure rail to another. At the same time, no complex control mechanism is needed to properly delay the valves as further discussed above. This simple and elegant architecture allows for immediate switching between the pressure rails (high-pressure rail to medium-pressure rail; medium-pressure to low-pressure rail; high-pressure to low-pressure; medium-pressure rail to high-pressure rail; and medium-pressure rail to low-pressure rail) without any cross-talk or short-circuit between the rails, while still granting independent pressure control in each one of the multi-chamber cylinder chambers, since the proportional valve provides a degree of pressure control downstream. Therefore, by adjusting the proportionality of the opening of the valve, fine-tune control is achievable given a supply and return rail selection. Second, only a single proportional valve is needed per chamber, instead of 2 or sometimes 3 like in the prior art. This approach results in a further advantage of lowering cost as well as control complexity.

According to one embodiment, a control scheme500for these three valve arrangements is shown inFIG.6. Although direct force control is also possible, the example is shown with an additional outer-loop controller502that can be used. The outer loop evaluates the difference between a reference signal for the state to be controlled and its actual measurement. Position (x) or speed (x) control are achieved with a PID controller, that adjusts the reference force command as shown inFIG.7. The gain Kpscales to error to create a control input proportional to error of the controlled state (speed in the example shown), while the gains KIand KDact on the tracking error integral and derivative respectively. All the three controller components are them summed to generate a force command, which is sent to a force mode selection algorithm.

The force mode selection algorithm receives the desired cylinder force, as well as the rails pressures and the cylinder speeds. It then selects the state of each on/off valves (uon/off) such that energy losses are minimized. A diagram of the algorithm is shown inFIGS.8aand8bwhich are two figures splitting the algorithms into two pages. With respect to the diagram and other figures provided herein, the variables provided therein are defined in Table-1, below.

TABLE 1Definition for variables used in figures of the present disclosureAA, ABand ACEffecctive areas of chambers A, B and Crespectively, with chambers A and Cacting in the positive direction and chamberB acting in the negative directionpch,A, pch,BandRequired Pressures in chambers A, B and Cpch,Cps,A, ps,BandSupply side pressure in proportional valvesps,Cconnected to chambers A, B and Cpr,A, pr,BandReturn side pressure in proportional valvespr,Cconnected to chambers A, B and CFrefReference Force CommandpmaxMaximum pressure allowed in the chamberpminMinium pressure allowed in the chamber{dot over (x)}ActuatorΔpA, ΔpBandPressure drop across proportional valvesΔpCconnected to chamber A, B and C respectively.FmodeForce that would be obtained at a given mode,when the pressure drop across the proportionalvalve is zero.iGeneric reference to different chamber.Can be either A, B or CModeMode number, in this case 1 ≤ Mode ≤ 27ModelastOptimal cylinder mode in previous time stepJIPenalty applied to infeasible modesJCEPenalty applied to mode transitions thatrequired large control efforttswTime elapsed since previous switchttargetTarget time for interval between switchestsController sampling timeJELPenalty to energy lossesJmodeTotal cost associated with a given modeJminTotal cost associated with the mode withminium costpref,A, pref,BReference pressure commands to chambersand pref,CA, B and C respectively

For each available mode, the code evaluates the

where JELis a penalty on energy losses, while JCEpenalizes the needed control effort for a switch, by avoiding frequent switches and J, penalizes modes that are not feasible in the current operating condition. The algorithm evaluated Jmodefor each of the modes available. InFIG.8b, the selection of the optimal solution is carried out in section 4 (block identified as “4”). In addition, since in each mode the pressures ps,A, ps,B, ps,Cand pr,A, pr,B, pr,Care defined, mode feasibility can be verified by evaluating the pressure differentials needed and achievable across the proportional valve, as highlighted in section 2 of the embodiment. In section 3, the controller penalizes mode switches when the time passed after the previous switch (tsw) is lower than a target time interval for the next switch (ttarget). In case constraints or targets are not met, penalties represented by large values (LV1) and (LV2) are used to penalized prohibited modes, therefore avoiding their selection.

The block receives actuator speed measurement which is utilized to evaluate the required amount of throttling losses in each mode. This is carried out by evaluating

where tsis the controller sampling time, and Fmodeis the resultant cylinder output force that would be available in case no proportional valve was used.

Additionally, the algorithm also evaluates the necessary pressures in each cylinder chamber such that Frefis achieved, as highlighted in section 1. This results in a reference pressure (pref,i) to each cylinder chamber.

Each cylinder chamber has their own local controllers with respective pressures being controlled by means of feedback control as shown inFIG.9. These controllers also receive information about the commanded states to the on/off valves (uon/off) such that the controller knows in advance the pressure in the supply and in return side of the proportional valve. Therefore, based on the received pressure levels at the high-pressure rail (php), medium pressure rail (pmp) and low pressure rail (plp) and on the valves status, the pressure evaluation logic block defines values for the pressure on the supply side (ps,i) and on the return side (pr,i) of the proportional valve. In this way, the opening of each proportional valve can be electronically compensated when there is a change in pressure either on the supply or in the return side of the proportional valve with a nonlinear valve map that evaluates the necessary valve command such that the desired flowrate is achieved at a given pressure differential. Such pressure differential (Δp) is obtained by evaluating the difference between the measured chamber pressure (pch,i,meas) and the pressure on the supply or return sides. This is necessary because both the supply side pressure and the return side pressure vary depending on the states of the on/off valves. The embodiment also shows a PID controller adjusts the flow command (Qcmd) to each chamber based on the values of the controller gains KP,i, KI,iand KD,isuch that the reference pressure is tracked. This flow command is used as input to the valve nonlinear map, which outputs a command to the respective proportional valve (upv,i)

Additionally, a novel approach in hydraulic architecture for heavy machinery is presented that can provide a large number of discretized forces without requiring a complicated actuator design. This allows the introduction of small throttle control for fine control adjustments through the proportional valves without a significant increase in the system losses and without the need for a cylinder with a high number of chambers, which can significantly increase cylinder design complexity and impact its reliability. Towards this end, reference is made toFIG.10which is a schematic of a hydraulic arrangement600for a heavy machinery system including an actuator601with three chambers and three pressure rails602,604, and606resulting in 27 discretized force levels. The hydraulic arrangement600inFIG.10is shown with one actuator601, although more-linear or rotary actuators-could be coupled to the same pressure rails. The linear actuator601includes three chambers independently controlled by a network of proportional valves (individually not identified for sake of simplicity). Each valve of the network of the proportional valves is responsible for coupling a chamber with one of the actuators to a supply rail. Three rails602,604, and606are provided, each one at a different pressure level.

These different pressure rails are generated by a power source, such as the one shown inFIG.10(internal combustion engine (ICE), or by other power generation schemes (e.g., electrical motor, etc.) known to a person having ordinary skill in the art. The pressure in each rail is controlled to remain within pre-determined limits defined according to the specific application, and hydraulic accumulators may be used in these lines. Even though, the embodiment represents the system with two variable displacement pumps supplying flow to the pressure rails, it is also possible to develop different configurations for the flow supply to the rails. These can include the use of fixed displacement pumps, or other variations known to those having ordinary skill in the area. With respect to the pressure control in the rails, feedback signals from one of more actuators position, speed, acceleration and/or force can be used to adjust the rails pressure range as well as to vary the flow from a hydraulic flow supply, e.g., a hydrostatic pump. In the shown architecture, therefore, there are 9 proportional valves, which can be either pilot or direct operated. They are coupled to the multi-chamber actuator601connecting each chamber to at least two of all the three supply rails602,604, and606. When all chambers are coupled to all three rails, 27 discrete levels of forces (33) are achieved, and within each level of the 27 discretized levels, a number of micro-adjustments afforded by the proportional valves is made possible. This arrangement provides a superior level of discretization over the prior art by significantly reducing the complexity of the cylinder designs from, e.g., 5 chambers to 3, while yielding a higher number of discrete force availability, therefore increasing the actuator efficiency. In addition, the higher system efficiency yielded by the higher number of force levels available allows the introduction of small throttle control in any of the three chambers for precise motion control while maintaining a high overall system efficiency. In terms of control approach, this architecture has similar structure to that already described inFIG.6,FIG.7andFIGS.8aand8b., with the only difference being the local pressure controller. A top level control block diagram is presented inFIG.11, with the details of the pressure controller being presented inFIG.12. In this case, a different valve selection logic is implemented, that selects the valve able to provide the desired flowrate (Qcmd) at the lowest pressure drop possible. This block informs the nonlinear valve maps which valves are active and which valves are not (uoff). The maps then output valve commands to each one of the proportional valves connected to chamber i. Consequently, the main difference is that, in this case, the controller outputs commands to three proportional valves-which are connected to the same chamber-being one connected to the HP pressure rail (upv,i,hp), one connected to the medium pressure rail (upv,i,mp) and one connected to the low pressure rail (upv,i,lp).