Variable delivery external gear machine

An external gear machine (EGM) includes a housing, an inlet, a drive gear positioned in the housing and configured to be (i) driven by a mechanism when the EGM is operated as a pump, or (ii) drive an external mechanism when the EGM is operated as a motor, the drive gear having a plurality of teeth, a slave gear positioned in the housing having a plurality of teeth and configured to be driven by the drive gear, an outlet formed in the housing and configured to receive at least some of the volume of fluid via an outlet fluid communication channel, a first slider defining an inlet fluid communication channel and the outlet fluid communication channel, selective positioning of the first slider configured to vary net operational volumes of fluid communication between the inlet and the outlet, for a given rotational speed of the drive gear.

TECHNICAL FIELD

The present application relates to gear machines, and specifically to external gear machines used in fluid power management systems.

BACKGROUND

External gear machines (EGMs) are used as primary flow supply unit in many applications such as fuel injection systems, small mobile applications such as micro-excavators, turf, and gardening machines. EGMs are also used in fixed applications such as hydraulic presses and forming machines. EGMs also find applications in auxiliary systems such as hydraulic power steering, fan drive systems and as charge pump in hydrostatic transmissions.

Referring toFIG. 1A, a perspective view of an example of an EGM10is depicted. The EGM10includes a housing12, a drive gear14, which drives a slave gear16, both disposed inside the housing12. The drive gear14and the slave gear16are supported by bushings18inside the housing12. The drive gear14and the slave gear16are coupled together in a mesh zone where a plurality of their respective teeth comes into contact with each other. Tooth space volumes between any two consecutive teeth of the drive gear14and any two consecutive teeth of the slave gear16pick up fluid and deliver fluid as the teeth rotate about the housing12. Specifically, in the mesh zone the tooth space volumes initially decrease as the respective teeth come into contact with each other and increase as the teeth come apart from each other. As the tooth space volume decreases, fluid pressure increases, causing ejection of fluid through an outlet22at an output pressure. Similarly, as the volume increases, the pressure decreases causing suction of fluid from the inlet24at an inlet pressure. End caps26and28enclose the housing12, where the end cap26provides a journal support for the drive gear14.

Referring toFIG. 1B, a perspective view of an example of the bushing18is provided. Fluid is communicated via an outlet fluid communication channel in the form of a groove30from the varying spaces between the teeth in the mesh zone to the outlet; and similarly fluid is communicated via an inlet fluid communication channel in the form of a groove32to the varying spaces between the teeth in the mesh zone from the inlet. Therefore, grooves permit to utilize the full volumetric capacity of the unit, avoiding localized pressure peaks and fluid cavitation. In a pressure compensated EGM, such as the one represented inFIG. 1A, these grooves are realized in the bushings18(FIG. 1B)—popularly known by the names of bearing blocks and pressure/thrust plates—used to realize optimal sealing of the tooth space volumes through a proper lubricating fluid film even at high operating pressures. In non-compensated EGM, these grooves are machined in the pump housing.

The above described principle of operation of an EGM makes these units inherently fixed displacement. This inability of adapting the fluid displaced per every revolution on the basis of user's requests makes EGMs unsuitable for applications in energy efficient system layout configurations which characterize many fluid power applications. In these system configurations, variable displacement units can offer energy saving even greater than 50% compared to solutions based on fixed displacement units.

These factors have driven significant research towards the definition of a working concept for variable displacement EGMs. Past effort can be broadly categorized into two different sets of solutions: the first set of solutions consists of changing the meshing length of the gears. Several patent references describe different solutions for this idea, by moving the gears axially (US2001024618, EP0478514, US2008044308, and US2002104313). The second set of solutions consists in changing the inter-axis distance between the gears, thereby affecting the meshing area of the gears as provided in at least two patent references (CN85109203 and GB968998). However, each of these solutions introduces significant technological challenges, such as complexity, and has not resulted in successful commercialization. In fact, several major issues have to be faced to implement a viable and cost effective solution to move the gears, which are the most mechanically loaded parts of the machine, requiring at the same time good sealing and smooth transmission of power between the gears.

Efforts to obtain variable flow supply units were also made at system level; in particular, solutions that combine fixed displacement pumps with fast switching valves controlled in pulse width modulation (PWM) to obtain a variable output flow were proposed by several researchers. Despite the theoretical validity of these so called “virtually variable displacement” solutions, their application in real systems is hampered by the limited time response of electromechanical valves as well as compatibility issues of current fixed displacement pumps with the introduction of severe pressure pulsations.

There is, therefore an unmet need for a novel approach to provide variable flow at low and high pressures in gear pumps.

SUMMARY

An external gear machine (EGM) is disclosed. The EGM includes a housing, an inlet formed in the housing and configured to receive fluid from a supply, a drive gear positioned in the housing and configured to be (i) driven by a mechanism when the EGM is operated as a pump, or (ii) drive an external mechanism when the EGM is operated as a motor, the drive gear having a plurality of teeth. The EGM further includes a slave gear positioned in the housing having a plurality of teeth and configured to be driven by the drive gear, the drive gear configured to engage the slave gear in an angular mesh zone, tooth space volumes defined by tooth spaces between each two consecutive teeth of the drive gear and each two consecutive teeth of the slave gear configured to receive volumes of fluid from the inlet via an inlet fluid communication channel as the corresponding teeth rotate about the inlet. In addition, the EGM includes an outlet formed in the housing and configured to receive at least some of the volume of fluid via an outlet fluid communication channel when the corresponding tooth space volumes in the angular mesh zone decrease as the corresponding teeth of the drive gear and slave gear come into contact with each other. The EGM also includes a first slider defining the inlet fluid communication channel and the outlet fluid communication channel, selective positioning of the first slider configured to vary net operational volumes of fluid communication between the inlet and the outlet, for a given rotational speed of the drive gear.

According to one embodiment, the teeth of the EGM are asymmetrical.

According to one embodiment, the asymmetry of each tooth is defined by a first angle between a first face of the tooth in relationship with a first radial line and by a second angle between a second face of the tooth in relationship with a second radial line.

According to one embodiment, the ratio of the first angle to the second angle is between about 1 and 1.81.

According to one embodiment, the EGM further includes a second slider (also having grooves similar to those in the first slider) defining a secondary inlet fluid communication channel and a secondary outlet fluid communication channel such that selective positioning of the second slider provides fluid cooperation with the inlet fluid communication channel and the outlet fluid communication channel in order to vary net operational volumes of fluid communication between the inlet and the outlet, for a given rotational speed of the drive gear.

According to one embodiment, the second slider and the first slider are operatively coupled to each other.

According to one embodiment, the first slider is operated by an electromechanical actuator.

According to one embodiment, the electromechanical actuator is a stepper motor.

According to one embodiment, the electromechanical actuator is a solenoid.

According to one embodiment, the first slider is operated by a mechanical actuator configured to move the first slider based on one of (i) pressure differential between the inlet and the outlet, (ii), pressure at the outlet, and (iii) a combination thereof.

A hydraulic displacement system (HDS) is also disclosed. The HDS includes a mechanism for (i) driving an external gear machine (EGM) when the EGM is configured to be a pump, or (ii) being driven by the EGM when the EGM is configured to be a motor. The HDS also includes a fluid supply. The HDS also includes an EGM. The EGM includes a housing, an inlet formed in the housing and configured to receive fluid from a supply, a drive gear positioned in the housing and configured to be (i) driven by the mechanism when the EGM is operated as a pump, or (ii) drive the mechanism when the EGM is operated as a motor, the drive gear having a plurality of teeth, a slave gear positioned in the housing having a plurality of teeth and configured to be driven by the drive gear, the drive gear configured to engage the slave gear in an angular mesh zone, tooth space volumes defined by tooth spaces between each two consecutive teeth of the drive gear and each two consecutive teeth of the slave gear configured to receive volumes of fluid from the inlet via an inlet fluid communication channel as the corresponding teeth rotate about the inlet, an outlet formed in the housing and configured to receive at least some of the volume of fluid via an outlet fluid communication channel when the corresponding tooth space volumes in the angular mesh zone decrease as the corresponding teeth of the drive gear and slave gear come into contact with each other and a first slider defining the inlet fluid communication channel and the outlet fluid communication channel, selective positioning of the first slider configured to vary net operational volumes of fluid communication between the inlet and the outlet, for a given rotational speed of the drive gear.

According to one embodiment, the teeth are asymmetrical.

According to one embodiment, the asymmetry of each tooth is defined by a first angle between a first face of the tooth in relationship with a first radial line and by a second angle between a second face of the tooth in relationship with a second radial line.

According to one embodiment, the ratio of the first angle to the second angle is between about 1 and 1.81.

According to one embodiment, the EGM further includes a second slider defining a secondary inlet fluid communication channel and a secondary outlet fluid communication channel such that selective positioning of the second slider provides fluid cooperation with the inlet fluid communication channel and the outlet fluid communication channel in order to vary net operational volumes of fluid communication between the inlet and the outlet, for a given rotational speed of the drive gear.

According to one embodiment, the second slider and the first slider are operatively coupled to each other.

According to one embodiment, the first slider is operated by an electromechanical actuator.

According to one embodiment, the electromechanical actuator is a stepper motor.

According to one embodiment, the electromechanical actuator is a solenoid.

According to one embodiment, the first slider is operated by a mechanical actuator configured to move the first slider based on one of (i) pressure differential between the inlet and the outlet, (ii), pressure at the outlet, and (iii) a combination thereof

The attached drawings are for purposes of illustration and are not necessarily to scale.

DETAILED DESCRIPTION

A novel approach for varying flow rate through an external gear machine (EGM), formed as a pump or motor, is described in the present disclosure. The external gear machine according to the present disclosure provides variable timing for fluid transfer from an inlet to an outlet of the machine. The described solution preserves the compactness, reliability and low cost features typical of an EGM and achieves control of flow displaced by the machine. The novel design concept further takes advantage of asymmetric involute and trochoid profiles of gears, which are used to maximize the range of flow variation achievable by the machine. The proposed design is also optimized to maximize the performance levels, in terms of delivery flow pulsations—typically responsible for noise emissions and vibrations—volumetric efficiency, internal pressure peaks and cavitation onset which occur during the meshing process of the gear of the EGM.

Referring toFIG. 2, an exploded perspective view of an EGM100, according to the present disclosure, is presented. The EGM100includes a housing112, a drive gear114, which drives a slave gear116, both disposed inside the housing112. The drive gear114and the slave gear116are supported by bushings118A and118B inside the housing112. The drive gear114and the slave gear116are coupled together in a mesh zone (depicted inFIG. 3) where a plurality of their respective teeth comes into contact with each other. End caps126and128enclose the housing112and coupled to the housing by fasteners119, where the end cap126provides a journal support127for endshaft115of the drive gear114. The EGM100also includes an outlet122and an inlet124. The EGM also includes end caps126and128,

The EGM100also includes sliders120A and120B. These sliders120A and120B are coupled to the respective bushings118A and118B. A sealing member is fastened to the housing120. The positioning and coupling of the sliders120A and120B with respect to the bushings118A and118B is described below with referenceFIG. 3.

Referring toFIG. 3, a plane view of the drive gear114and the slave gear116in engagement with each other is provided. The drive gear114has a plurality of teeth, exemplified by202A and202B, while the slave gear116also has a plurality of teeth, exemplified by204A and204B. Tooth space volume206is identified as the space between any two consecutive teeth. Within this space, fluid is picked up and then trapped between any two consecutive teeth of the drive gear114and any two consecutive teeth of the slave gear116and the housing112. The engagement of the teeth creates a mesh zone210identified as the angular portion θ. It should be noted that the tooth space volume206is a variable that is constant for most of its rotational path but begins to decrease and then increase within the mesh zone210.

The mesh zone is divided into four portions. The first portion (identified as 1 in a circle) is the upper portion inFIG. 3, where the teeth just begin to engage each other. This portion is identified as the space between mesh-zone-start214A and upper-exterior-portion216A. As the teeth from both the drive gear114and the slave gear116come together in the first portion of the mesh zone (1), the space volumes206of the respective gears begin to interfere with each other and the overall tooth space volumes206decrease. As the tooth space volumes206decrease, fluid pressure increases, causing ejection of fluid through the outlet122at an output pressure. At this point fluid begins to be ejected from the EGM100via an outlet grove222(also referred to as the outlet fluid communication channel), identified in dashed lines for clarity, positioned below the mesh zone210as well as openings (not shown) to the outlet122. The bottom of the first portion is identified by the point “D” which signifies a point in the rotation where the teeth have trapped the fluid in the associated tooth space volumes206as a result of contact with each other. Beyond point “D” the only path for ejection of fluid is through the outlet groove222to the outlet122. In other words, point “D” corresponds to the switch point between i) fluid ejection via the outlet groove222and other openings (not shown) to ii) fluid ejection via the outlet groove222only by isolating tooth space volumes206with the outlet groove222.

The second portion (identified as 2 in a circle) is the upper-interior portion inFIG. 3. This portion is identified as the space between the upper-exterior-portion216A and the centerline218. As the tooth space volume decreases, fluid pressure increases. In this portion the teeth come in contact with each other and trap the fluid within the shrinking tooth space volume206. Somewhere in this portion (2), the outlet groove ends, at which point fluid is no longer able to be ejected via the outlet groove222. At the center212of mesh zone210the tooth space volumes206are minimized. At any point beyond the center212, the tooth space volume206begins to increase.

The third portion (identified as 3 in a circle) is the lower-interior portion inFIG. 3. This portion is identified as the space between the centerline218and lower-exterior-portion216B. In this portion the teeth remain in contact with each other and continue to trap the fluid, however, now the tooth space volumes206begin to increase. Somewhere in this portion (3), an inlet groove224(also referred to as the inlet fluid communication channel), shown in dashed lines for clarity, ends; at which point fluid that is isolated to the inlet groove224can begin to be sucked in via the inlet groove224from the inlet124. The end of portion3is designated as “S” inFIG. 3, corresponding to a switch point between i) fluid suction via the inlet groove224only by isolating tooth space volumes206with the inlet groove224to ii) fluid suction via the inlet groove224and other openings (not shown) to the inlet124.

The fourth portion (identified as 4 in a circle) is the lower portion inFIG. 3, where the teeth just begin to separate from each other. This portion is identified as the space between lower-exterior-portion216B and mesh-zone-end214B. As the teeth from both the drive gear114and the slave gear116come apart from each other in the fourth portion of the mesh zone (4), the space volumes206of the respective gears continue to expand. As the tooth space volumes206increase, the fluid pressure decreases causing suction of fluid from the inlet124at an inlet pressure. At this point fluid continues to be sucked into the EGM100via the inlet grove224positioned below the mesh zone210as well as openings (not shown) to the inlet124.

A no-fluid-communication-zone226is depicted between the bottom of the outlet groove222and the top of the inlet groove224. This zone226corresponds to an angular space in which fluid is not communicated to either the inlet or the outlet. Minimizing this zone226, maximizes fluid displacement, however, too much of this zone226, can cause pressure spikes and cavitation resulting in noise and other mechanical issues.

Referring toFIG. 4, a graph of tooth space volume vs. angular position for the arrangement depicted inFIG. 3is provided. The aforementioned four portions of the mesh zone210are marked in successive numbering from 1 to 4. In portions1and2, fluid is ejected from the EGM100. In portions3and4fluid is sucked into the EGM100. The lowest point of the graph (marked as M), represents a switching point where fluid is no longer ejected but begins to be sucked in. The point M coincides with the centerline218and the center212inFIG. 3.

Referring toFIG. 5, a gear interface300is presented that according to the present disclosure provides for a different fluid transfer than what is provided inFIG. 3. There are two main differences between the gear arrangements inFIGS. 3 and 5, firstly the disposition of the outlet groove222vs.322in which the outlet groove322(also referred to as the outlet fluid communication channel) is elongated into the no-fluid-communication-zone226. Secondly, the inlet groove324differs from224, wherein inlet groove324(also referred to as the inlet fluid communication channel) is shortened away from the no fluid communication-zone226. The elongation of the outlet groove322as well as the simultaneous shortening of the inlet groove324, results in a reduced flow delivery (equivalent to a lower pump displacement). This flow reduction is because during the portion2of the mesh zone210and part of the portion3of the mesh zone210, each tooth space volume206remains coupled to the outlet port122via the outlet groove322while the tooth space volumes206decrease and then while the tooth space volumes206begin to expand. As a result a part of the fluid already delivered to the outlet is taken back into the tooth space volumes206via the outlet groove322which acts as a “fluid dead volume.” This “dead volume” in effect varies the net operational volume that passes through the EGM. Net operational volume is defined as the net volume of fluid sucked into the EGM from the inlet124and the volume of fluid ejected to the outlet122(considering the fluid dead volume). Graphically, the principle can be represented by a larger (as compared toFIG. 4) portion coupled to the outlet. The additional “dead volume” is equal to the difference between the volumes of points S and M, therefore the effective fluid displaced to the outlet is equal to the difference between the maximum volume and volume at point S.

While the difference betweenFIGS. 3 and 5and the associated graphs provided inFIGS. 4 and 6have been based on elongation of the outlet grooves322as well as the shortening of the inlet grooves324, the same effect can be provided by providing the inlet groove and the outlet groove on sliders that can be moved with respect to the centerline218. The sliders120A and120B were shown inFIG. 2. The inlet and outlet grooves can be machined in these sliders120A and120B to provide a path for fluid communication to the outlet122and the inlet124. Referring toFIG. 7the sliders120A and120B are shown in their respective slots within the bushing118A and118B, respectively. Therefore, with respect toFIG. 3, the elongation of the outlet groove222to the outlet groove322along with the simultaneous shortening of the inlet groove224to the outlet groove324, shown inFIG. 5, can be achieved by sliding the sliders120A and120B downward towards the inlet124. As seen inFIG. 7, each slider120A and120B includes an outlet groove120A1and120B1, respectively, that are configured to couple tooth space volumes to the outlet122; and an inlet groove120A2, and120B2, respectively, which are configured to couple the tooth space volumes to the inlet124. Each slider120A and120B also includes a switch zone121A and121B, respectively, configured to switch between ejecting fluid to the outlet122and sucking fluid from the inlet124.

To realize delivery flow variation, the slider can move either towards the inlet port124or towards the outlet port122. However, being the conditions of the fluid at the inlet port is often close to saturation (for the case of a pump); it is preferable to consider the motion towards the inlet port, so that cavitation effects due to fluid aeration are limited. The opposite consideration applies for the case of a motor (a distinction between motors and pumps will be made further in the present disclosure).

It should be appreciated how the variation of the achieved flow delivery as a result of change in slider position of sliders120A and120B can occur for all slider positions that realize the described switching between the points D and S, in which each tooth space volume is trapped between points of contact of the teeth. By moving the slider outside the limits D-S a direct bypass connection between the outlet and the inlet would be realized, hence significantly reducing the volumetric efficiency of the variable delivery (VD)-EGM. With symmetric gears, the points D and S lie very closely to point M and hence they do not offer a large variation in the displacement, therefore novel asymmetric gear profiles are used to maximize the reduction in displacement.

Gears with asymmetric teeth profile, unconventional for EGMs, were investigated with the particular aim of maximizing the range of displacement variation achievable for the (VD)-EGM. The design of the teeth includes involute and trochoid profiles above and below the base circle, respectively. In order to accomplish the goal of designing asymmetric teeth, two different pressure angles are considered respectively for the drive and opposite coast tooth flanks as shown inFIG. 8A. In order to ensure that the asymmetrical teeth gear profile is physically manufacturable using conventional manufacturing processes such as hobbing, shaping, rack-cutting etc., an asymmetrical cutter profile is assumed at first. The tooth profile is then derived on the basis of the shape of the asymmetric cutter as shown inFIGS. 8A and 8B.

Based on the design variables, the parameters which govern the shape of the asymmetric cutter are obtained using Eqs. (1)-(4),

Having the analytical tool to develop the asymmetric teeth, we now turn our attention to the inlet and outlet groove profiles. In the VD-EGM, the grooves machined in the lateral bushings (or in the housing, for not pressure compensated designs) perform the important timing function of connecting tooth space volumes (TSVs) with the inlet or outlet environment when the TSV is trapped between points of contact. Therefore, they contribute in determining the amount of fluid displaced per revolution by every TSV. With an optimal crossport (simultaneous connection of the TSV with the inlet and outlet port), the grooves can also ensure minimal internal pressure overshoots and localized cavitation effects during the transition of TSV from/to the low pressure and high pressure regions. For the asymmetric gear profile, a particular “two-winged”structure of the grooves was developed as provided inFIG. 9A. The different parameters which control the shape of the grooves are depicted inFIGS. 9A and 9B. The main intent for using such a two winged architecture with four angular controls (a's) is to control the influence of the pressure angles (drive and coast side) of the gear profiles on the performance of the machine. Particular emphasis is placed on the feasibility of machining the grooves using the conventional milling process for prototyping. As can be seen fromFIGS. 9A and 9B, the radius ‘R’ of the milling tool is taken into consideration, so that the results of the optimization process can be directly prototyped without any additional consideration based on manufacturability. It should be noted that the groove profiles displayed in theFIGS. 9A and 9bare not the only designs which are applicable to the design of VD-EGM. Other groove profiles can be used as long as they are machined on a movable slider.

The maximum reduction in flow delivery (also referred as minimum displacement condition) can be calculated by investigating the location of the points D and S (which define the angular locations at which the fluid in the TSV is trapped between the contact points between the gears) in the curve that characterize the volume of each TSV, as shown inFIGS. 10A and 10B. Due the asymmetric nature of the gears, the location of the points ‘D’ and ‘S’ are not symmetric about the point ‘M’ (angular location which the trapped TSV is at a minimum) unlike gears with symmetric involute teeth. It can be seen however that the angular range remains at the same point (θD2−θS2=θD1−θS1) though the actual position at which these points occur differs as seen inFIGS. 10A and 10B,

where θD2is the angular location at which the fluid in the tooth space volume begins to be trapped between the points of contact of the slave gear,

θS2is the angular location at which the fluid in the tooth space volume seizes to be trapped between the points of contact of the slave gear,

θD1is the angular location at which the fluid in the tooth space volume begins to be trapped between the points of contact of the drive gear, and

θS1is the angular location at which the fluid in the tooth space volume seizes to be trapped between the points of contact of the drive gear. Since dual flank configuration is imposed on all the gears, and in order to expand further the angular range of the trapped volume, both the drive TSV and slave TSV behave as two independent displacing chambers which are not connected to each other. Therefore, to maximize the full potential in achieving the reduction in displacement, the switch of the connection of the drive TSV from the delivery to suction should occur at point S1, and at point S2for the TSV on gear2since both TSVs operate as separate displacement chambers due to the introduction of dual flank configuration.

The resultant minimum displacement achievable can be expressed as an average of the ones provided by the drive and the slave TSVs independently. The minimum displacement achievable can be calculated using Eq. (11)

β=βdrive+βdriven2=VS⁢⁢2+VS⁢⁢12·VM(11)
where βdriveis minimum displacement % of the drive tooth space volume,
βdriven(Slave)is minimum displacement % of the slave tooth space volume, la
Vs2is volume of the slave tooth space volume at Point S,
Vs1is volume of the drive tooth space volume at Point S, and
VMis volume of the minimum tooth space volume.

Experimental results of various configurations are provided with reference toFIGS. 11A and 11B. It can be seen fromFIGS. 11A and 11Bthat, according to at least one embodiment the flow rate proportionally (68%) reduces at minimum displacement condition as compared to those at full or maximum displacement, however other reduction amounts may be possible depending on gear diameter specification. A good agreement between simulated data and measurements can be observed from these figures. However, at both displacements there is a small offset between the two curves which can be explained by the low tolerance of the process used to realize the gears. The gears do not permit to strictly maintain the dual flank contact conditions for all teeth into mesh; as a consequence, an imperfection in achieving zero backlash between the gears is introduced, and a certain amount of bypass leakage is introduced from the high pressure to the low pressure side through the TSVs hence causing a lower volumetric performance in the experiments.

FIGS. 12A and 12Brepresent the input shaft torque validation. It can be seen that the input shaft torque reduces proportionally at minimum displacement. In one embodiment, approximately 32% reduction in torque is obtained at all the operating conditions tested for minimum displacement.

The validations for volumetric efficiency are provided inFIGS. 13A and 13B. The curve with diamond dots represents the performance of the reference commercial EGM (of the same displacement). It can be noticed that the volumetric performance at maximum displacement, as shown by the curve with square dots matches very closely to that of the reference design (shown by the curve with diamond dots). It should also be noted that the volumetric efficiency for all the pressures at maximum and minimum displacement is higher for 2000 rpm (FIG. 13B) as compared to 1000 rpm (FIG. 13A), this is because the maximum speed at which the casing was broken in was 2000 rpm and hence the gears are capable of achieving better radial sealing thereby leading to better volumetric performance at 2000 rpm.

As expected, it can be seen fromFIGS. 13A and 13Bthat the volumetric efficiency at minimum displacement is lower than the volumetric efficiency at maximum displacement. This is due to the internal leakages (at the tip of the teeth and in the lateral side of both gears) are most prominently dependent on the pressure and hence have a larger influence for the efficiency at lower displacement. The trends of simulated volumetric efficiency at minimum displacement matches closely to that of the measured values.

Returning toFIG. 2, EGM100according to the present disclosure can be operated as a pump, wherein the inlet124is coupled to a low pressure supply of fluid (not shown) and the outlet122is coupled to a downstream apparatus (not shown) that is being actuated by the pump, and where the drive gear114is driven by an external actuator (e.g., electric motor, internal combustion engine etc.) (not shown). Alternatively, the EGM100can be operated as a motor, wherein the inlet124is coupled to a high pressure supply of fluid and the outlet122is coupled to a fluid reservoir (not shown) where the drive gear114drives an external apparatus (not shown).

Furthermore, the sliders120A and120B can be actuated by a pressure differential apparatus that uses pressure differential at the outlet122and the inlet124to position the sliders to maintain a desired volume displacement. Alternatively, the sliders120A and120B can be operated by an actuator such as a stepper motor that is controlled by a controller via a processor that senses outlet pressure at the outlet122and inlet pressure at the inlet124and adjusts the position of the sliders120A and120B accordingly.

The EGM according to the present disclosure can also be operated in a system, where the EGM is either operated as a pump or a motor. Referring toFIG. 14, such a block diagram of such a system400is depicted. The system400includes a mechanism402coupled to the drive gear (similar to the drive gear114ofFIG. 2) of an EGM404, the mechanism is configured to (i) drive the EGM404when the EGM404is configured as a pump, or (ii) be driven by EGM404, when the EGM404is configured to be operated as a motor. Therefore, the connection between the mechanism402and the EGM404is provided in dashed format. The EGM400is fluidly coupled to a fluid supply406. The EGM404is also optionally fluidly or mechanically coupled to a load actuator408configured to facilitate actuation of a load. The load actuator408may optionally be coupled to the fluid supply406.