Patent Publication Number: US-9410544-B2

Title: Charged hydraulic system

Description:
CROSS REFERENCE TO RELATED APPLICATION 
     This application is a continuation-in-part of U.S. application Ser. No. 12/261,195, filed Oct. 30, 2008, which claims foreign priority benefits under U.S.C. §119 from European Patent Application No. 07254336.6 filed on Nov. 1, 2007, each of which is hereby incorporated by reference in its entirety. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates to hydraulic systems with at least one hydraulic high-pressure pump and at least one hydraulic charging pump according to the generic part of claim  1 . Furthermore, the invention relates to hydraulic pumps. 
     BACKGROUND OF THE INVENTION 
     Hydraulic systems are nowadays used for a plethora of different purposes. 
     One prominent example is the use of hydraulics for generating large forces. For this purpose, usually cylinders and pistons are used. Such devices are used, for example, in locks, steering systems, crawlers, forklift trucks, wheel loaders, and so on. Hydraulic systems for these types of machines are usually referred to as open-circuit hydraulics. This notation is used, because within the hydraulic actuator, for example in the hydraulic cylinder, a variable volume of hydraulic fluid is present. To compensate for these volume changes, a hydraulic fluid reservoir is provided. The hydraulic fluid reservoir is under atmospheric pressure and is usually built as a standard tank. To perform its function as a buffer for the hydraulic fluid, the tank usually has to be of considerable size. Since the hydraulic fluid in the reservoir is under atmospheric pressure, the hydraulic pump takes in hydraulic fluid directly from an atmospheric fluid reservoir. This is a main difference between open-circuit hydraulic systems and closed-circuit hydraulic systems, which are described in the following. 
     Another application where hydraulic components became very popular are transmissions for vehicles which benefit from continuous variable ratio and wheelspeed combined with high tractive effort over the whole speed range and especially at low speeds. Such transmissions very often use closed-circuit hydraulic pumps and closed-circuit hydraulic motors. The hydraulic motor converts the high-pressure energy of the hydraulic fluid into mechanical energy and sends the hydraulic fluid, now at a lower pressure level, back to the hydraulic pump. Such a system is generally referred to as closed-circuit hydraulics, because the hydraulic pump is sending and receiving almost the same flow rate of hydraulic fluid under all working conditions of the hydraulic circuit. Therefore, no buffer is needed. The low pressure side of such systems normally operates between 10 and 30 bars. Because of this closed-circuit systems normally have fewer problems with filling of the hydraulic pump than open-circuit hydraulic systems. 
     In real applications, however, even a closed-circuit hydraulic system still has some hydraulic fluid reservoir under atmospheric conditions. First of all, leakage of hydraulic fluid has to be considered. Especially in devices with mechanically moving parts, such as in hydraulic pumps and hydraulic motors, fluid leaks can never be totally avoided. The leakage fluid is therefore collected and transferred to the fluid reservoir via collecting lines. The collected hydraulic fluid is pumped back into the closed-circuit hydraulic system (normally to the low-pressure side of the circuit) by means of a charge pump. Sometimes, a small fraction of hydraulic fluid is taken out of the closed hydraulic circuit for cooling and filtration purposes. This is commonly referred to as “loop flushing”. A pressure relief valve and/or an orifice take out a certain percentage of the total fluid flow rate on the low pressure side of the closed-circuit hydraulic system. This flush part of the fluid flows through a heat exchanger and heat can be transferred from the hydraulic fluid to the ambient air. Having passed the heat exchanger and optionally a fluid filter, the fluid is ejected to the hydraulic fluid reservoir. From there, it is pumped back to the main fluid circuit by means of a charge pump, together with the leakage hydraulic fluid. The fraction of hydraulic fluid, used for cooling and filtration purposes, is relatively small and is lower than about 20 percent of the fluid flow rate in the main hydraulic circuit. 
     While hydraulic systems perform well in practice, they are still undesirably large and expensive for certain applications. 
     Especially in open-circuit hydraulic systems, problems arise in high performance conditions. Under such high performance conditions the hydraulic pump has to deliver a large flow rate of hydraulic fluid. This, of course, requires the hydraulic pump to receive an appropriate amount of hydraulic fluid from the fluid reservoir. To be able to do this, the suction line of the hydraulic fluid pump has to have a huge cross section, so that a sufficient fluid supply rate to the hydraulic fluid pump can be provided and the pressure drop can be kept low. However, not only the suction line has to have a large cross section, but also the fluid inlet port (e.g. the valve plate of an axial piston machine) of the hydraulic pump needs to be designed with a sufficiently large cross-section. These requirements for large supply cross sections result in relatively large sizes of pump and motor parts, fittings, flanges, hoses and pipes and hence of the overall size of the resulting hydraulic system. This leads to increased costs for the manufacture and use of such hydraulic systems, especially when considering the increased volume requirements in the machine or vehicle, where the hydraulic system is used. 
     In check ball pump designs the inlet check valve always means an additional flow restriction and the aforementioned problem increases. Normally this results in limited fill speed of such pumps. Very often the inlet valve is actually held close by a spring and the fluid has to work against the spring. The pump has to suck the inlet valve open. Synthetically commutated hydraulic pumps are very similar to check ball pumps when considering the aforementioned problem. In such synthetically commutated hydraulic pumps, also known as digital displacement pumps (which are a unique subset of variable displacement pumps), the fluid valves do not open passively under the influence of pressure differences. Instead, the fluid valves are actively controllable by appropriate valve actuating units which are controlled by an electronic control unit. Even when the inlet valve in a synthetically commutated hydraulic pump is of the normally open type, it provides additional inlet flow restriction which limits fill speed when the pump takes in hydraulic fluid from an atmospheric hydraulic fluid reservoir. 
     These synthetically commutated hydraulic pumps fall into two groups. In the first group, only the inlet valve is actively controlled, whereas the fluid outlet valve remains passive. With this type, a full stroke pumping mode, a partial stroke pumping mode and a no-pumping mode can be obtained. With the second type, where both inlet and outlet valves are of the actively controllable type, a full or partial stroke back pumping mode/motoring mode can be realised as well. This is known in the state of the art. 
     The requirement of a large supply cross-section is a major drawback for synthetically commutated hydraulic pumps. Not only valve cross-sections, and therefore the valve head in the valve channel, have to be of large size, but also the valve actuating unit has to be able to deliver a sufficiently large force as well as a sufficiently large travel. This, in turn, increases the costs for such a hydraulic pump. Moreover, the driving unit of the valve has high power consumption. This increases the costs for the manufacture and the actual use of such a hydraulic system even further. On off-highway mobile equipment for instance this would require the installation of large and expensive alternators to generate sufficient electrical power for inlet valve actuation. 
     SUMMARY OF THE INVENTION 
     The object of the invention is therefore to provide a hydraulic system with an increased overall performance. Another object of the invention is to provide a hydraulic pump with an increased overall performance. 
     A hydraulic system and a hydraulic pump, showing the features of the respective independent claims, solve the problem. 
     It is suggested, that a hydraulic system with at least one hydraulic high-pressure pump and at least one hydraulic charging pump, in which the output hydraulic fluid flow of said hydraulic charging pump is used as the input hydraulic fluid flow of said hydraulic high-pressure pump is designed in a way, that the maximum flow rate of said output fluid flow of said hydraulic charging pump is at least 50 percent of the maximum flow rate of said input fluid flow of said hydraulic high-pressure pump. Put in other words, the performance of the hydraulic charging pump is chosen in a way that it can provide a sufficiently high fluid flow rate, so that this fluid flow rate together with the fluid flow rate being returned from the hydraulic consumers, is sufficiently high, to provide the hydraulic high-pressure pump with a sufficiently high input fluid flow rate, so that the hydraulic high-pressure pump can be running at full speed and maximum displacement, at least under all working conditions which normally can be expected. This, of course, should be even true, if the hydraulic system is an open-circuit hydraulic system, where only a relatively small amount of hydraulic fluid or no hydraulic fluid at all is returned to the input port of the hydraulic high-pressure pump (at least not directly). As long as these conditions are met, the actual percentage can defer from 50 percent as well. For instance, 30 percent, 40 percent, 60 percent, 70 percent, 80 percent and/or 90 percent could be used as a percentage. 
     Using the suggested design, the pressure of the hydraulic fluid on the fluid supply side of the hydraulic high-pressure pump is elevated above ambient pressure. Therefore, even with the same supply cross section, the fluid supply can be increased, as compared to standard, uncharged hydraulic high-pressure pumps. Therefore, it is possible to decrease the size of the supply cross sections, to increase the performance of the hydraulic high-pressure pump, and/or to increase the maximum shaft speed and/or pumping flow rate of the hydraulic high-pressure pump. If the hydraulic high-pressure pump is of the synthetically commutated type, it is also possible to decrease the power consumption of the pump. Particularly it is possible to decrease the electrical power consumption of the actuated valves (if electrical power is used for valve actuation). Further advantages are, that the proposed hydraulic system can be used at higher altitudes and, because of the decreased risk of cavitation, the wear of the hydraulic high-pressure pump can be decreased. 
     Preferably, the maximum flow rate of said output fluid flow of said hydraulic charging pump is at least essentially the same as or higher than the maximum flow rate of said input fluid flow of said hydraulic high-pressure pump. With this design, it is possible to run the hydraulic system at high performance levels even in situations, where no hydraulic fluid at all (at least not directly) is returned from the hydraulic consumer. This design is particularly useful in open circuit hydraulic systems, of course. In particular, the maximum flow rate of said output fluid flow of said hydraulic charge pump can be 100 percent, 105 percent, 110 percent, 115 percent, 120 percent, 125 percent or 130 percent of the maximum flow rate of said input fluid flow of said hydraulic high-pressure pump. This way, leakages can be accounted for and the loop flushing principle can be implemented. 
     The output pressure of said hydraulic charging pump can be regulated to be between 0.3 to 10 bars, preferably 0.5 to 7 bars, more preferably 1 to 5 bars, even more preferably 1.5 to 3 bars, most preferably 2 to 2.5 bars. The given pressures are meant to be pressures above ambient atmospheric pressure (or standard atmospheric pressure). Even a slight increase in the charging pressure of the hydraulic high-pressure pump can lead to a significant increase in performance. This can be easily understood, when considering a pressure drop of 0.3 bars along the fluid supply line (including the fluid inlet valve) as an example: If the fluid reservoir has a pressure, which is equal to the atmospheric pressure, the pressure drop amounts to 30 percent of the pressure available. If, however, the input-pressure is charged to 1 bar above atmospheric pressure (i.e. 2 bars absolute) the pressure drop is now only 15 percent of the total pressure available. Roughly speaking, this can lead to a performance increase of about 50 percent. Because a quite small pressure increase by the charging pump is sufficient, the loading pump can be quite small, simply and durably designed and inexpensive to manufacture. Nevertheless, the overall performance can be increased substantially. 
     If necessary, a plurality of hydraulic high-pressure pumps and/or a plurality of hydraulic charging pumps can be provided. It is possible, that a single hydraulic charging pump supplies several hydraulic high-pressure pumps. On the contrary, it is also possible that a plurality of hydraulic charging pumps serve a single hydraulic high-pressure pump. Also, it is possible that several pumps are arranged in parallel, wherein every hydraulic high-pressure pump has its own, dedicated hydraulic charging pump. 
     In a preferred embodiment of the invention, at least one hydraulic high-pressure pump is a synthetically commutated hydraulic pump. As already mentioned, the proposed hydraulic system is particularly useful when synthetically commutated hydraulic pumps are used. Although it is possible that the hydraulic charging pump is of a synthetically commutated type as well, normally a different type of pump is chosen for the hydraulic charging pump for cost reasons. In general, synthetically commutated hydraulic pumps, particularly charged synthetically commutated hydraulic high-pressure pumps have the following advantages: They have smaller and cost effective inlet (flow pressure) valves; they have a higher flow speed, even at high or maximum displacement of the pump; they have smaller ports and smaller diameters of supply lines (e.g. hoses, pipes and fittings); they can have smaller internal ports and hence reduction in size and weight is possible; prevention of cavitation and hence less wear is possible; the hydraulic system can be used at higher altitudes. 
     It is suggested that at least two hydraulic pumps are driven by the same power source. Especially, a hydraulic high-pressure pump and its dedicated hydraulic charging pump can be driven by the same power source. As a power source, a combustion engine, an electric motor, a turbine or the like can be used. In particular, a power source could mean a mechanical power source. The power source can be connected to the pumps by a rotatable shaft, for example. 
     Preferably, at least one hydraulic charging pump is of a self-delimiting type. By a self-delimiting type, a design is meant, wherein a pressure increase on the output side of the pump automatically delimits the fluid flow rate, pumped by the change pump. For example, an impeller-like pump can be used. 
     Also, instead of a self-delimiting pump, a pump, in particular a positive displacement pump, could be used as a charge pump in which a check valve or a pressure relief valve is used to purge excess flow back from the charging pump to the hydraulic fluid reservoir. Such a circuit can have similar performance like the use of a “genuine” self-delimiting charge pump. Such a purge valve can also be useful, when several flow sources are combined for charging, e.g. flow from the charge pump, return flow from the main system (driven by the hydraulic high-pressure pump) and/or return flow from another sub-system (e.g. a steering system supplied with hydraulic fluid by a separate hydraulic pump, e.g. a gear pump). These different flow sources might be decoupled from each other by additional check valves, if necessary. The check valve with appropriate spring rate can purge excess flow back to the reservoir tank and can ensure that sufficient charge pressure at the right level will be available. In cases where synthetically commutated hydraulic high-pressure pumps are used as high-pressure pumps, the purge valve can also allow flow reversal through the hydraulic high-pressure pump during motoring mode. 
     In particular, it is suggested that at least one hydraulic charging pump is of a fluid jet pump type. The design is based on the principle of a water ejector pump. This design can be very simple, durable, inexpensive and self-delimiting. As the driving fluid jet, the hydraulic fluid, being returned from a hydraulic consumer, or the fluid flow of a special pump can be used. Particularly in off-highway applications, very often a second pump is used to provide flow to another sub-system. A typical sub-system can be a steering system supplied e.g. by a gear pump as the second pump. The return flow from such a sub-system (e.g. from the steering system) can be used to drive the fluid-jet pump. 
     Preferably, at least one hydraulic pump is designed as a two stage pump. Particularly a hydraulic high-pressure pump is designed as a two stage pump. Using such a design, it is possible to design the pumps very simple and inexpensive. Such an integrated two stage pump can be especially suitable for systems with one dedicated charge pump per hydraulic high-pressure pump. Nevertheless, a relatively high overall charging pressure and/or flow rate can be provided for the hydraulic high-pressure part of the pump. An example is the use of a fluid-jet type pump or an impeller type pump as a charging stage. In particular, such a two-stage pump can be used as the only pump, present in the hydraulic system. Also, a charging pump of the system can be a two-stage pump as well. For example, an impeller pump could drive a fluid jet pump. 
     A possible embodiment of the invention can be obtained when the output fluid flow of the hydraulic high-pressure pump is joined with the output fluid flow of the hydraulic charging pump, after the output fluid flow of the hydraulic high-pressure pump has passed a hydraulic consumer, and the thus combined fluid flows are used as the input fluid flow of the hydraulic high-pressure pump. Here, the still somewhat elevated pressure of the hydraulic fluid, even after the hydraulic fluid has passed the respective hydraulic consumer, can be used as a charged input fluid flow. The elevated pressure can even be created artificially by inserting a check valve with an appropriate spring rate. This can save energy, because it is not necessary to first reduce hydraulic fluid pressure to ambient pressure and to pressurise the hydraulic fluid again. If a high capacity charging pump is used, the high-pressure pump—and therefore the whole hydraulic system, including the hydraulic consumer, supplied by the fluid flow of the high-pressure pump—can still run at full performance, even in conditions, where not all flow from the hydraulic system or consumer (or even only a minor fraction of the flow, pumped to the hydraulic system or consumer) is returned because of e.g. the use of differential hydraulic cylinders. 
     Preferably, the output fluid flow of at least one hydraulic charging pump is used at least partially for a hydraulic consumer. Partially can stand for a mode, where the output fluid flow rate of the hydraulic charging pump is used for a hydraulic consumer during certain time intervals. Alternatively or additionally, it is possible that a certain fraction of the output fluid flow rate of the hydraulic charging pump is used for a hydraulic consumer. The hydraulic consumer can be a device with low priority, or at least with a lower priority than the hydraulic consumer, which is supplied by the hydraulic high-pressure pump. For instance, the output of the hydraulic high-pressure pump could be used for a steering device, while the low priority consumer is a mixing device of a concrete delivery truck. By such a design, the hydraulic charging pump can be used in an optimal manner. 
     Another possible embodiment of the invention can be achieved, if at least one hydraulic consumer can be alternatively supplied by the output fluid flow of at least one hydraulic high-pressure pump and/or the output fluid flow of at least one hydraulic charging pump. This design is particularly useful for a hydraulic consumer that can be run at several pressure levels, whereas certain functions or a certain output force of the hydraulic consumer can only be reached at higher pressures. If, for instance, the hydraulic consumer is a hydraulic cylinder for lifting loads, the hydraulic cylinder can be fed by the charging pump, if only small loads are to be moved. However, the speed can be high, due to the high output-fluid flow rate of the charging pump. Also, energy can be saved. If, however, heavy loads are to be lifted, the hydraulic cylinder can be moved by the hydraulic high-pressure pump, although the speed is slower. 
     A very compact and preferable design of a hydraulic pump can be achieved, if the hydraulic pump comprises at least a first, charging stage and a second, high pressure stage. By such a design, a hydraulic charging pump and a hydraulic high-pressure pump can be integrated into just one device. This device can be used as a drop-in solution for already existing hydraulic systems. 
     Preferably, the charging stage can comprise an impeller device and/or a fluid jet device. Using such a design, the already mentioned effects and advantages can be achieved for a two-stage hydraulic pump in a similar way, as well. 
     Preferably, both stages are driven by a common driving shaft, and are preferably mounted on said driving shaft. This design is particularly useful, if an impeller pump is used. Once again, the already described advantages and effects can be achieved similarly. 
     Another embodiment of the invention can be achieved, if the output hydraulic fluid flow of the hydraulic charging pump is at least partially going through a hydraulic consumer, before being used as the input fluid flow of the hydraulic high-pressure pump. This aspect of the invention can even be used in conventional closed circuit hydraulic systems, particularly in closed circuit systems with a loop flushing. By the proposed design, the energy output of the hydraulic charging pump can be used, for instance, during operation modes where a lower output flow rate of the hydraulic charging pump is needed, and the performance of the charging pump can therefore be used for generating a higher pressure, instead of generating a higher fluid flow rate. By this design, already mentioned effects and advantages can be achieved in a similar way. 
     Although in the previous description, as well as in the following description, references are made mainly to hydraulic pumps, it is to be understood, that the hydraulic pumps can also be used in a reversed pumping mode and/or a motoring mode, as well. However, the proposed invention, as well as its suggested various designs are particularly useful in the full and/or part-stroke pumping mode. 
     If, however, the hydraulic high-pressure pump should be used in a motoring mode, it is possible to by-pass the charging pump, using a check valve with an appropriate spring rate, for example. It is also possible to use both pumps in a motoring mode, of course. Another possibility is, that the charging pump is of a design, so that it is essentially no problem for the respective pump, when fluid flow is reversed. Fluid jet pumps can, for instance, be of such a design. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       The objects, advantages and effects of the present invention will be elucidated by the following description of certain embodiments of the invention, which are described using the enclosed figures. The figures are showing: 
         FIG. 1  is a schematic diagram of a first example of a charged hydraulic circuit, wherein a single charging pump and a single high-pressure pump are used; 
         FIG. 2  is a schematic diagram of a second example of a charged hydraulic circuit, wherein a two-stage charging pump and a single high-pressure pump are used; 
         FIG. 3  is a schematic diagram of a third example of a charged hydraulic circuit, wherein the hydraulic circuit is an only partially open circuit hydraulic system; 
         FIG. 4  is a schematic diagram of a fourth example of a charged hydraulic circuit, wherein the return flow of a hydraulic consumer is used to drive a jet pump, which is used as the charge pump; 
         FIG. 5  is a schematic diagram of a fifth example of a charged hydraulic circuit, wherein several high-pressure pumps and several hydraulic consumers are present and which is an only partially open circuit hydraulic system; 
         FIG. 6A  is a first example of an integrated hydraulic pump with a charging stage and a high-pressure stage; 
         FIG. 6B  is a second example of an integrated hydraulic pump with a charging stage and a high-pressure stage; 
         FIG. 7  is a schematic cross section through a synthetically commutated hydraulic pump; 
         FIGS. 8A, 8B  is an illustration of the mutual dependency of the different fluid flow rates in charged hydraulic systems; and 
         FIG. 9  is an exemplary example, illustrating the principles, shown in  FIG. 8A /B. 
     
    
    
     DETAILED DESCRIPTION 
     In the following description, the same reference numbers are used for similar devices, shown within different figures. This does not necessarily mean, that the referenced devices are identical in design or function. However, the principle function or design of the respective device is similar. 
     In the figures one common drive shaft  11  for all pumps is shown. Of course the pumps can also be driven by different shafts and with different shaft speeds. This is often the case when some pumps are driven by the crank shaft of a combustion engine and some other pumps are e.g. mounted on a PTO (Power Take Off; split drive shaft) of the engine or the gear box. In such cases the different shaft speeds have to be considered during system design. However, this does not limit the applicability of the invention. 
       FIG. 1  shows a schematic diagram of a charged, open-circuit hydraulics  1 . The hydraulic circuit  1  comprises a charging pump  2 , a synthetically commutated hydraulic pump  3  (also known as digital displacement pump or variable displacement pump), serving as a high-pressure pump, a hydraulic machine  4 , powered by the pressurised hydraulic fluid and a fluid tank  5 , serving as a reservoir for the hydraulic fluid. The components are interconnected by fluid lines  6 ,  7 ,  8 ,  9 ,  60 , which may be hoses, pipes or internal passages within an assembly. 
     The charging pump  2  and the synthetically commutated hydraulic pump  3  are driven by a common mechanical energy source  10 , in the example shown a combustion engine, via a common rotatable shaft  11 . Therefore, whenever the combustion engine  10  is running, both the charging pump  2  and the synthetically commutated hydraulic pump  3  are driven at the same time. 
     Although not shown, the combustion engine  10  can also drive an electric generator, producing electric energy, which can be used for powering the actively controlled valves of the synthetically commutated hydraulic pump  3 . 
     The hydraulic machine is of a type, where the input fluid flow, provided by the high-pressure line  8 , is not necessarily equal to the hydraulic output fluid flow to the returning line  9 . For example, the hydraulic machine  4  could be a hydraulic cylinder. Therefore, the volume of hydraulic fluid within the hydraulic circuit  1  is highly variable. Excess charge flow from charge pump  2  which is not needed by high-pressure pump  3  is purged via charge pressure relief valve  18  and pressure relief line  60  back to the fluid tank  5 . The pressure relief valve  18  is of course only needed when charge pump  2  is of a non-self-delimiting type, e.g. a positive displacement type. 
     To compensate for these variations in “captured” hydraulic fluid volume, a sufficiently large fluid tank  5 , containing hydraulic fluid, is provided. The fluid tank  5  is exposed to ambient pressure, i.e. usually about one bar. However, in certain applications, such as in planes or in machinery, designed to be used at high altitudes (e.g. mountainous areas) this pressure can be much lower. 
     The hydraulic fluid, contained within the fluid tank  5 , is sucked into the charging pump  2  via suction line  6 . To minimise the pressure losses between the fluid tank  5  and the charging pump  2 , and to maximise the fluid throughput, the suction line  6  and the inlet area of the charging pump  2  show relatively large cross sections. The charging pump  2  pressurises the hydraulic fluid to a slightly elevated pressure, which is present in the mid-pressure line  7 , and adjacent parts of the charging pump  2  and the synthetically commutated hydraulic pump  3 . In the example, shown in  FIG. 1 , the elevated pressure is chosen to be about 2 to 3 bars above ambient pressure. 
     Although the pressure difference between ambient pressure and elevated pressure is relatively low, the increase in performance of the hydraulic circuit  1  is quite remarkable. Because of the elevated pressure within the mid-pressure line  7 , the mid-pressure line&#39;s  7  cross section can be smaller, and still a high fluid flux can be achieved. 
     More important, however, not only the cross section of the mid-pressure line  7 , but also the cross sections of the fluid inlet line  54  and the inlet valves fluid cross sections  57  can be chosen smaller, and still a sufficient fluid flow rate can be maintained (see  FIG. 7 ). Also, the speed of the synthetically commutated hydraulic pump  46  can be chosen higher, because of the higher input fluid flow (this idea can be used for other circuits as well). 
     The hydraulic fluid, pressurised by the synthetically commutated hydraulic pump  3 , is expelled into the high-pressure line  8 . Typical pressure values for the high-pressure line  8  are between 200 bars to 500 bars, depending on the application. However, different pressures can be chosen as well. 
     The high-pressure line  8  is connected to the hydraulic machine  4 , thus providing the hydraulic machine  4  with the necessary fluid supply rate. The fluid machine  4  can be almost any suitable hydraulic machine, known in the state of the art. A detailed description is omitted for brevity. 
     Finally, the hydraulic fluid, leaving the hydraulic machine at a reduced pressure, is returned to the fluid tank  5  via the returning line  9 . 
     In  FIG. 2 , an example for a two-stage charged, open-circuit hydraulics  16  is shown. 
     Similar to the open circuit hydraulics  1 , shown in  FIG. 1 , the two-stage charged hydraulic circuit  16  according to the example shown in  FIG. 2 , comprises a charging pump  2 , a synthetically commutated hydraulic pump  3 , a hydraulic machine  4  and a fluid tank  5 . Charging pump  2  and synthetically commutated hydraulic pump  3  are driven by combustion engine  10  via a common rotatable shaft  11 . 
     Contrary to the open circuit hydraulics  1 , shown in  FIG. 1 , in the present example of a two-stage charged hydraulic circuit  16 , the output fluid flow of the charging pump  2  is not going directly to the synthetically commutated hydraulic pump  3 , but instead the output fluid flow is directed through the elevated pressure line  22  to a second charging pump  12 , which is designed as a fluid jet pump  12  in the example shown. The basic design of fluid jet pump  12  is similar to a hydrostatic jet pump, used e.g. in chemistry. Therefore, the hydraulic fluid, entering the fluid jet pump  12  through the elevated pressure line  22 , will cause additional hydraulic fluid, to be sucked in from the fluid tank  5  into the fluid jet pump  12  through the second suction line  15 . Therefore, an “amplified” fluid flow will leave the fluid jet pump  12  in the direction of the mid-pressure line  14 . The mid-pressure line  14  will feed the synthetically commutated hydraulic pump  3 , which in turn will feed the hydraulic machine  4 . 
     The fluid jet pump  12  converts the pressure energy of the hydraulic fluid in the elevated pressure line  22  into an increased amount of hydraulic fluid at the lower pressure level of the mid-pressure line  14 . A comparatively small and inexpensive charging pump  2  can therefore provide a quite large fluid flow rate for the synthetically commutated hydraulic pump  2 , with the help of the fluid jet pump  12 . 
       FIG. 3  shows an example for a partially closed circuit hydraulics  17 . Once again, the partially closed circuit hydraulics  17  comprises a synthetically commutated hydraulic pump  3  and a charging pump  2 , which are driven by a combustion engine  10  via a common rotatable shaft  11 . 
     The hydraulic circuit  17 , shown in  FIG. 3 , is partially closed, in the sense that the fluid flow, leaving the synthetically commutated hydraulic pump  3  in the direction of a first hydraulic machine  19  via the high-pressure line  8 , is not necessarily returned to the fluid reservoir  5  after leaving the first hydraulic machine  19 . Instead, the fluid, leaving the first hydraulic machine  19 , enters the mid-pressure line  14  which serves as the fluid input line for the synthetically commutated hydraulic pump  3 . However, the partially closed circuit hydraulics  17  still differs from normal closed circuit hydraulics, and even from a closed circuit hydraulics using a loop flushing, as will be come clear from the following description. 
     In the partially closed circuit hydraulics  17 , the first hydraulic machine  19  can be of a type where the input fluid flow and the output fluid flow of said first hydraulic machine  19  can be substantially different. So the first hydraulic machine  19  can be in a working condition, where the return fluid flow is substantially higher (e.g. twice as high) as the input fluid flow. It is even possible that the first hydraulic machine  19  does not receive any hydraulic fluid at all, but does return a substantive amount of hydraulic fluid. In such condition the hydraulic fluid entering the mid-pressure line  14  exceeds the amount of hydraulic fluid, leaving the mid-pressure line  14  through the synthetically commutated hydraulic pump  3 . This excess amount will be discharged by a spring loaded check valve  18  into the fluid tank  5  through returning line  9 . 
     If, on the contrary, the first hydraulic machine  19  uses hydraulic fluid, without returning any hydraulic fluid into the circuit (or returning only a small fraction of the input fluid flow rate), the hydraulic fluid now needed in the mid-pressure line  14  will be provided through the charging pump  2 . The charging pump  2  accepts hydraulic fluid from the fluid tank  5  via the suction line  6  and will discharge this hydraulic fluid at an elevated pressure into the elevated pressure line  13 . Before entering the mid-pressure line  14 , the hydraulic fluid first performs some useful work in the second hydraulic machine  20 . It should be noted that the charging pump  2  is able to pump hydraulic fluid and therefore to power the second hydraulic machine  20  in any working state of the partially closed circuit hydraulics  17  or first hydraulic machine  19 , because excess fluid in the mid-pressure line  14  will be discharged through the spring loaded check valve  18  into the fluid tank  5 . 
     The partially closed circuit hydraulics  17  can be equally realised if the second hydraulic machine  20  is omitted and replaced by a simple fluid line. Also, a bypass-line, bypassing the second hydraulic machine  20  at least in part, can be provided. 
     It should be understood that the exact pressure levels of the high pressure line  8 , the elevated pressure line  13 , the mid-pressure line  14 , the suction line  6  and the return line  9  might be different from the respective line, shown in the examples of  FIGS. 1 and 2 . This statement is true for all figures. 
     In  FIG. 4 , a schematic diagram of a modified partially closed circuit hydraulics  21  is shown. In some sense, the modified partially closed circuit hydraulics is a combination of ideas, taken from  FIG. 2  and  FIG. 3 . 
     The modified partially closed circuit hydraulics  21  again comprises a charging pump  2  and a synthetically commutated hydraulic pump  3 . Both pumps are driven by a combustion engine  10  through a common rotatable shaft  11 . 
     The fluid, expelled by the synthetically commutated hydraulic pump  3  is fed to the first hydraulic machine  19  via the high-pressure line  8 . Hydraulic fluid, leaving the first hydraulic machine (where the ratio of the input flow rate and output flow rate can vary) is returned directly to the fluid tank  5  via the returning line  9 . However, the input fluid flow of the synthetically commutated hydraulic pump  3  does not come directly from the charging pump  2  (via a direct line, a bypass-line or via the second hydraulic machine  20 ). 
     Instead, the hydraulic fluid is sucked in by the charging pump  2  from the fluid tank  5  via suction line  6  and expelled to the elevated pressure line  13 . From there, the hydraulic fluid performs some work in the second hydraulic machine  20  from where it is expelled into the connecting line  22 . This fluid flow is used as a driving input of a fluid jet pump  12 . As already described, the fluid jet pump  12  “amplifies” the fluid flow, flowing through the stage connecting line  22 , and the thus “amplified” common fluid flow is expelled into mid-pressure line  14 . The mid-pressure line  14  serves as the input line for the synthetically commutated hydraulic pump  3 . Spring-loaded check valve  18  (or alternatively a pressure release valve) is used as a purge valve to spill excess charge flow from mid-pressure line  14  via return line  9  to fluid tank  5 . Since charge pump  12  is of a self delimiting type in this example, purge valve  18  is optional and not essential for the protection of the charge pump  12  and for the hydraulic system. However, the spring-loaded check valve  18  would be necessary, if the charge pump  12  is constructed in a way that no “backward flow” from connecting line  22  to second suction line  15  is possible. Of course, a bypass-line, bypassing the second hydraulic machine  20  can be provided as well. 
     Of course, such a spring loaded check valve  18  can be used at different places and within different embodiments, as well. For instance, such a spring loaded check valve  18  could be used in the example of  FIG. 2  between elevated pressure line  22  and return line  9  and/or between mid-pressure line  14  and return line  9 . However, if in the examples of  FIGS. 1 and 2  the charging pumps  2  are of a self-limiting type, such a spring-loaded check valve  18  can be omitted as well. 
     In  FIG. 5 , a multi machine hydraulic circuit  23  is shown as another example of a hydraulic circuit. To some extent, the multi machine hydraulic circuit  23  of  FIG. 5 , resembles the partially closed circuit hydraulics  17  of  FIG. 3 . 
     Hydraulic fluid from the fluid tank  5  enters the charging pump  2  via suction line  6 . 
     The multi machine hydraulic circuit  23  comprises a single charging pump  2  and three synthetically commutated hydraulic pumps  3   a ,  3   b ,  3   c , which are driven by the same combustion engine through a rotatable shaft  11 . 
     The hydraulic fluid expelled by the charging pump  2  enters the second hydraulic machine  20  via the elevated pressure line  13 . The hydraulic fluid, leaving the second hydraulic machine  20  (or bypassing the second hydraulic machine  20  via a bypassing line) forms part of the fluid flow, entering the mid-pressure line  14 , which is the feeding line for the synthetically commutated hydraulic pumps  3   a ,  3   b ,  3   c . In case there is an excess flux into the mid-pressure line  14 , a spring loaded check valve  18  serves as a relief valve and hydraulic fluid is expelled to the fluid tank via returning line  9 . 
     The high-pressure output of the three synthetically commutated hydraulic pumps  3   a ,  3   b ,  3   c  is expelled into respective high pressure lines  8   a ,  8   b ,  8   c . First hydraulic machine  19  and third hydraulic machine  24  are directly connected with first high pressure line  8   a  and third high pressure line  8   c , respectively. 
     Additionally, three electrically actuated valves  26   a ,  26   b ,  26   c  are provided. Using first electrically actuated valve  26   a , first high pressure line  8   a  and second high pressure line  8   b  can be fluidly connected or disconnected. Similarly, using second electrically actuated valve  26   b , second high pressure line  8   b  and third high pressure line  8   c  can be fluidly connected or disconnected. 
     Using third electrically actuated valve  26   c , it is possible to connect second high pressure line  8   b  to elevated pressure line  13 , and therefore to second hydraulic machine  20 . A check valve  25  is provided between second high pressure line  8   b  and elevated pressure line  13  for safety reasons. In case consumer  20  is a steering system, check valve  25  assures that at least the output flow from pump  2  is exclusively available for consumer  20 . 
     By appropriately switching the electrically actuated valves  26   a ,  26   b ,  26   c , an optimum performance of the multi machine hydraulic circuit  23  can be reached for almost every thinkable workload condition of the three hydraulic machines  19 ,  20 ,  24 . 
       FIG. 6A  shows a first example of a dual stage hydraulic pump  27 , comprising a charging stage  28  and a high pressure stage  29 . The dual stage hydraulic pump therefore integrates a charging pump  2  and a synthetically commutated hydraulic pump  3  into a single pump  27 . Both stages  28 ,  29  are driven by a common rotatable shaft  30 . 
     Hydraulic fluid, entering the synthetically commutated dual stage hydraulic pump  27  through a fluid inlet  31  with a large fluid supply cross section  32 , first reaches the charging stage  28  of the synthetically commutated dual stage hydraulic pump  27 . The charging stage  28  is essentially comprised of a plate  33  and an impeller disc  34 , which is arranged adjacent to the plate  33 . When the shaft  30  is turning, hydraulic fluid is pumped to mid-pressure chamber  35 . Here, the hydraulic fluid rests at an elevated pressure of 2 or 3 bars above ambient pressure, for example. The high pressure stage  29  of the synthetically commutated dual stage hydraulic pump  27  comprises pistons  40 , turnably sliding on a wobble plate  41 . When the shaft  30  is rotated, the wobble plate  41  causes the pistons  40  to reciprocally move in and out of their respective cylinder spaces  42 . Thus, a working chamber  37  of cyclically changing volume is provided. In a pumping mode, when the volume of the working chamber  37  increases, the inlet valve  36  (which is electrically actuatable) will be opened by an appropriate actuator unit. Because of the pressure present in the mid-pressure chamber  35 , the hydraulic fluid is not only sucked into the working chamber  37  by under-pressure within the working chamber  37 , but is also pushed into the working chamber  37  by the pressure within the mid-pressure chamber  35 . Because of this, the fluid supply cross-section of the inlet valve  36  can be smaller, compared to common hydraulic pumps. Furthermore, higher operating speeds of the synthetically commutated dual stage hydraulic pump  27  can be reached. Is should be noted, that in the example shown, a higher driving speed will lead to a better performance of the loading stage  28  as well, so that the pressure in the mid-pressure chamber  25  will increase accordingly. 
     As soon as the volume of the working chamber decreases, inlet valve  36  will be closed (at least in the full stroke pumping mode) and passive outlet valve  38  will open, as soon as an appropriate pressure difference between the working chamber  37  and the high pressure fluid line  43  has been established. 
     However, it is still possible to switch the synthetically commutated dual stage hydraulic pump  27  to a partial stroke pumping mode. The elevated pressure in the mid-pressure chamber  35  is not that high, that fluid cannot be expelled back into the mid-pressure chamber  35  from the working chamber  37 . 
     The high-pressure fluid lines  43  of the synthetically commutated dual stage hydraulic pump  27  connect within the pump&#39;s body to a common fluid manifold  44 . The fluid manifold  44  is consequently connected to a fluid output port  45 . 
       FIG. 6B  shows a second example of a dual-stage hydraulic pump  60 , comprising a charging stage  28  and a high-pressure stage  29 . Up to a quite large extent, the two examples of the dual-stage hydraulic pumps  27 ,  60  shown in  FIG. 6A  and  FIG. 6B , are similar to each other. Therefore, the same reference No. are used for similar parts. 
     In particular, the high-pressure stage  29  of the dual-stage hydraulic pump  60  is almost identical to the dual-stage hydraulic pump  27 , shown in  FIG. 6A . The details can therefore be looked up from the previous description. Different from the first example  27  in  FIG. 6A , the present dual stage hydraulic pump  60  of  FIG. 6B  shows a different charging stage  28 . In the present embodiment, the charging stage  28  shows a fluid jet pump  39 . As commonly known, a fluid jet pump  39  consists essentially of an injector  61  and a venturi channel  62 . In the present example, the entrance of the venturi channel  62  is fluidly connected to a fluid reservoir  5 . The injector  61  is fed by the return flow from a hydraulic consumer, e.g. by the return flow from a power steering. The pressure can be at 10 bar, while the flow rate can be set at 10 l/min. Using the fluid jet pump  39 , the fluid flow, flowing through the injector  61  is amplified by the flow, flowing through the venturi channel  62 , and the combined fluid flows (back flow from power steering and additional flow from a reservoir) are entering the mid-pressure chamber  35 . 
     Because of the charging stage  28  being designed as a fluid jet pump  39 , the plate  33  and the impeller disc  34 , which is present in  FIG. 6A , can be omitted. 
       FIG. 7  shows a standard synthetically commutated hydraulic pump  46 , as known in the state of the art. The cyclically changing working chamber  47  is formed by a piston part  48  and a cylinder part  49 . The cylinder part  49  and the piston part  48  are moved reciprocally in and out of each other by the joint forces of a cam  50 , mounted on a rotatable shaft  51  and a spring  52 , pushing the piston part  48  and the cylinder part  49  away from each other. An electrically actuated inlet valve  53  connects the inlet line  54  to the working chamber  47 . Accordingly, a fluid outlet valve  55  connects the working chamber  47  to a fluid outlet line  56 . 
     As can be seen from the standard synthetically commutated hydraulic pump  46 , shown in  FIG. 7 , the fluid supply cross-section  57  of the inlet valve  53  has to be very large. The valve head has to be very large. Therefore, a appropriately strong valve actuating unit  59  has to be provided. This valve actuating unit  59 , however, uses a lot of energy. 
     In  FIGS. 8A and 8B  a schematics of the different fluid flow rates in the vicinity of the hydraulic charge pump  2  and the hydraulic high-pressure pump  3  is shown. From this, conclusions about the sizing of the charge pump  2  and the high-pressure pump  3  can be drawn. 
     To prevent cavitation of the high-pressure pump  3  (which is preferably of the synthetically commutated type) the pressure on the inlet port  61  of the hydraulic high-pressure pump  3  has to be maintained at a suitable level under all operating conditions as already described earlier. To make the whole hydraulic pumping system of a certain machine as cost effective as possible, the charge pump  2  should be made as small as possible. If possible (which depends mainly on the hydraulic consumers) the output flow from the charge pump q cpout  (where  cpout  stands for “charge pump output flow rate”) and the return flows from the sub-systems q return  are combined and elevated to a suitable charge pressure using for instance the check valve  18  with a suitable spring rate. Alternatively a pressure relief valve or maybe even a correctly sized orifice can be used. To be able to sustain such a suitable charge pressure, the following equation should hold:
 
 q   return   +q   cpout   =q   hpin   +q   chexec   (1),
 
where q return  is the return flow rate from sub-systems, q cpout  is the charge pump output flow rate, q hpin  is the charge pump inlet flow rate and q chexec  is the excess charge flow rate, which is returned to the fluid tank  5 . Of course, in practice usually only positive values are possible for the different fluid flow rates.
 
     The exact value of the charge pressure at the inlet port  61  of the hydraulic high-pressure pump  3  might vary under different operating conditions but the system has to be designed in a way that under all circumstances sufficient charge pressure is provided and cavitation in the hydraulic high-pressure pump  3  is prevented. 
     If no return flow from sub-systems is available (i.e. q return =0) the charge pump has to be sized in a way that sufficient charge pressure for the hydraulic high pressure pump  3  is always guaranteed. In such a case a self-delimiting charge pump, e.g. an impeller or a jet pump, might be the most cost effective solution. In this case, a purge valve  18  can even be omitted, because equation (1) can be solved with a constant q chexec =0. This is because q cpout  will be automatically set to the appropriate level by the self-delimiting behaviour of charge pump  2 . 
     However, it is also possible to use a positive displacement pump for the charge pump  2 , together with a purge valve  18 . 
     It should be mentioned, that it is also possible to solve equation (1) by reducing q hpin . If in a hydraulic system at most only once in a while the fluid flow demand on the high-pressure side q hpout  is very high or the return flow rate from sub-systems q return  is very low, the pumping rate of the high-pressure pump  3  can be reduced by an electronic controlling unit (not shown). This way, cavitation in the high-pressure pump  3  can be avoided as well. Of course, the fluid output flow rate q hpout  will be correspondingly low. However, for certain applications this might not be a problem, especially if this situation only rarely occurs. 
     In  FIGS. 8A and 8B , two different basic designs of the hydraulic high-pressure pump  3  are illustrated. 
       FIG. 8A  shows a hydraulic high-pressure pump  3  with inlet port  61 , outlet port  62  and additional leakage collecting port  63 , to return internal leakage  64  to the fluid tank  5 . 
       FIG. 8B  shows a similar circuit that uses the hydraulic high-pressure pump  3  without a dedicated port for internal leakage  64 . 
     In  FIG. 8A  the high-pressure pump&#39;s input flow rate q hpin  has to make up for the oil flow on the leakage port  63  q hpleak  ( hpleak  for “high-pressure leakage”). This is not necessary for the system, shown in  FIG. 8B , because the internal leakage  64  of the hydraulic high-pressure pump  3  stays inside the hydraulic high-pressure pump  3  and does not have to be replaced. 
     The following equations can be used for charge pump sizing:
 
 q   hpout   +q   hpleak   =q   hpin   (2)
 
 q   hpin   +q   chexec   =q   return   +q   cpout   (3),
 
where q hpout  is the high-pressure pump output flow rate, q hpleak  is the high-pressure pump internal leakage flow rate, q hpin  is the high-pressure pump inlet flow rate, q chexec  is the excess charge flow rate returned to fluid tank  5 , q return  is the return flow rate from the sub-systems and q cpout  is the charge pump output flow rate.
 
     The system designer should ensure that always a minimum charge excess flow q chexec  remains through the purge valve  18 . The limit is when q chexec  becomes zero. In this case equation (3) becomes
 
 q   hpin   =q   return   +q   cpout   (4)
 
and
 
 q   hpout   +q   hpleak   =q   return   +q   cpout   (5).
 
     In case no return flow from hydraulic sub-systems is present (i.e. q return =zero) we will get
 
 q   hpout   +q   hpleak   =q   cpout , in case of FIG. 8A  (6)
 
 q   hpout   =q   cpout , in case of FIG. 8B  (7).
 
     The system designer should make sure that these rules are fulfilled under all operating conditions. In particular it is important to clearly understand return flow rates q return  from loads especially when differential hydraulic cylinders are involved. 
       FIG. 9  shows another example of a hydraulic system and how the return flows from several hydraulic consumers  19 ,  20  can be used in a cost effective manner for charging the hydraulic high-pressure pump  3   a . Pump  3   b  is a second hydraulic high-pressure pump. For cost reasons, most likely a fixed displacement pump will be used for second hydraulic high-pressure pump  3   b  (instead of a synthetically commutated hydraulic pump, as used for first hydraulic high-pressure pump  3   a ). Pump  3   b  acts as a supplement pump to supply extra flow on a high-pressure level directly into hydraulic consumer  19  if needed—e.g. for a higher propel speed of a vehicle, driven by a hydraulic motor. Switching of valve  26   a  will be synchronised with changing the output flow rate of synthetically commutated pump  3   a  by an electronic controlling unit (not shown). Since synthetically commutated pumps can change their output flow rate almost instantaneously, they can compensate switching supplement pump  3   b  in and out in an almost ideal manner. Particularly, the combined fluid output flow rate of first and second hydraulic high-pressure pumps  3   a  and  3   b  can be continuous. 
     As a guideline for the sizing of the pumps in particular for the sizing of the first and second hydraulic high-pressure pump  3   a ,  3   b , supplement high-pressure pump  3   b  ideally should be slightly smaller than first hydraulic high-pressure pump  3   a . This assumes, that both pumps  3   a ,  3   b  are driven at the same speed. Otherwise, the ratio of the different shaft speeds has to be considered for the design of the systems. For the present description, however, it is assumed that all pumps are driven with the identical shaft speed through a common shaft  11 . 
     Making supplement high-pressure pump  3   b  smaller than first hydraulic high-pressure pump  3   a  ensures that the high performance (high bandwidth) pump  3   a  maintains control of a flow rate, pressure etc. into hydraulic consumer  19 . 
     As soon as valve  26   a  activates high-pressure supplement pump  3   b  (flow from supplement pump  3   b  is added into hydraulic consumer  19 ) first high-pressure pump  3   a  has to instantaneously reduce its output flow rate to maintain constant input flow rate into hydraulic consumer  19 . 
     Because high-pressure supplement pump  3   b  is at least slightly smaller than first high-pressure pump  3   a  the return flow from hydraulic consumer  19  plus the flow from purge line  65  is not sufficient to charge the first high-pressure pump  3   a . In the embodiment shown in present  FIG. 9  the missing charge flow rate comes from a third pump  2  which like the high-pressure supplement pump  3  intakes hydraulic fluid from the atmospheric fluid reservoir  5  directly. The total displacement of pump  2  and high-pressure supplement pump  3   b  has to be at least equal to, but realistically bigger than the displacement of first high-pressure pump  3   a . How much bigger depends on the internal leakages and the type of the hydraulic consumer  19  used. In case hydraulic consumer  19  is a hydraulic motor (or several hydraulic motors in series or parallel) the return flow from hydraulic consumer  19  will be the input flow into hydraulic consumer  19  minus the leakage of the motors. In such case the total displacement of pump  2  and high-pressure supplement pump  3   b  only has to be slightly bigger than the displacement of first high-pressure pump  3   a . In case hydraulic consumer  19  contains differential cylinders or the like, the worst case (i.e. lowest ratio of input flow rate and return flow rate to and from hydraulic consumer  19 , respectively) has to be considered for sizing of pump  2 . In the same way the internal architecture of hydraulic consumer  20  has to be considered. In case hydraulic consumer  20  is a steering system the output flow rate of hydraulic consumer  20  should be very close to the input flow rate at all times (internal leakage of hydraulic consumer  20  is smaller). 
     Still referring to  FIG. 9 , the first hydraulic high-pressure pump  3   a  is charged by the high-pressure supplement pump  3   b  and the third pump  2 , as well as by itself through the outlet line of hydraulic consumer  19  since the fluid returned from both hydraulic consumer  19  and hydraulic consumer  20  are used to charge the first hydraulic high-pressure pump  3   a  through summation point  66 . Thus the high-pressure supplement pump  3   b  always charges the first hydraulic high-pressure pump  3   a , independently of the position of the valve  26   a . In particular, the high-pressure supplement pump  3   b  may charge the first hydraulic high-pressure pump  3   a  either directly through the valve  26   a  or by supplying supplemental flow to the hydraulic consumer  19  through the valve  26   a , which is then returned to the input of the first hydraulic high-pressure pump  3   a.    
     The system designer should make sure that under all operating conditions the total flow rate into summation point  66  is sufficiently high to provide suitable charge pressure into first high-pressure pump  3   a . If this can be guaranteed it might be better to choose one of the other proposed architectures and e.g. use a self-delimiting charge pump. One preferred case is a system in which the hydraulic consumer  19  are hydraulic motors and hydraulic consumer  20  a steering system. In this case high-pressure supplement pump  3   b  is switched in for higher road speeds. In this particular case the maximum power of the engine only allowed relatively moderate system pressures for higher road speeds and a gear pump for high-pressure supplement pump  3   b  was selected according to a certain exemplary embodiment. This resulted in a very cost effective overall system layout. 
     For example, in some embodiments, in which the hydraulic consumer  19  is one or more hydraulic motors and hydraulic consumer  20  is a steering system, the supplemental pump  3   b  may be a cheaper pump used to supplement flow into hydraulic consumer  19  when needed, such as for a higher propel speed as discussed above. One such exemplary embodiment may be, for example, in a typical forklift, where engine horsepower may only be sufficient for the high-pressure pump  3   a  to supply the consumer  19  with the necessary maximum pressure, e.g. a pressure increase of 420 bar, at low speeds (and low flow rates), but may not be powerful enough to provide the necessary maximum pressure at high flow rates. 
     In these embodiments, an advantage of providing the high-pressure pump  3   a  as a synthetically commutated pump may be utilized in connection with a cheaper fixed displacement pump as the supplemental pump  3   b  because the synthetically commutated pump is able to change displacement rapidly (much faster than any conventional pump with a servo mechanism). When the fixed displacement supplemental pump  3   b  is switched to provide flow directly to the hydraulic consumer  19 , i.e. using the valve  26   a , to supplement the synthetically commutated high-pressure pump  3   a , the synthetically commutated high-pressure pump  3   a  may, almost instantaneously, “jump down” in flow rate. Thus, after connecting the output of the fixed displacement supplemental pump  3   b  to the consumer  19 , the flow rate to the consumer  19  initially remains the same and, therefore, does not change the propel speed of the vehicle, thereby effectively providing a smooth transition. Then, as higher speeds are needed, the synthetically commutated high-pressure pump  3   a  may start ramping up flow rate, thereby allowing the relatively cheap fixed displacement pump (e.g. a pump having a low pressure rating) to supplement flow into the consumer  19  to provide the system with a high flow rate in addition to a high pressure. 
     Preferably, the fixed displacement supplemental pump  3   b  is smaller in displacement than the synthetically commutated high-pressure pump  3   a  so that the synthetically commutated high-pressure pump  3   a  still has headroom to decrease flow rate when the supplemental pump  3   b  is switched into the system (e.g. in case of engine stall when, for example, as the supplemental pump  3   b  is switched on, the vehicle starts to go uphill or hits an obstacle on the road or the like). For example, in some embodiments, the fixed displacement supplemental pump  3   b  may provide approximately 80% of the displacement of the synthetically commutated high-pressure pump  3   a.    
     Since the fixed displacement supplemental pump  3   b  is preferably smaller than the synthetically commutated high-pressure pump  3   a , the fixed displacement supplemental pump  3   b  may not be sufficient when used for directly charging the synthetically commutated high-pressure pump  3   a , as discussed above. Therefore, the third pump  2  may be configured to account for this difference. For example, in the exemplary embodiment where the fixed displacement supplemental pump  3   b  is 80% of the synthetically commutated high-pressure pump  3   a , the third pump  2  is preferably at least 20% of the synthetically commutated high-pressure pump  3   a  to account for the missing 20% flow. Thus, the combination of the third pump  2  and the supplemental pump  3   b  provides sufficient flow when the supplemental pump  3   b  is directly charging the synthetically commutated high-pressure pump  3   a.    
     While the present invention has been illustrated and described with respect to a particular embodiment thereof, it should be appreciated by those of ordinary skill in the art that various modifications to this invention may be made without departing from the spirit and scope of the present invention.