Technical Field
The following relates generally to an improved impact wrench, and more generally relates to an improved impact wrench having dynamically tuned drive components, such as an anvil socket combination and corresponding method of optimizing the characteristic functionality thereof.
State of the Art
Impact tools, such as an impact wrench, are well known in the art. An impact wrench is one in which an output shaft or anvil is struck by a rotating mass or hammer.
The output shaft is typically coupled to a fastener engaging element, such as a socket, configured to connect with a fastener (e.g. bolt, screw, nut, etc.) to be tightened or loosened, and each strike of the hammer on the anvil applies torque to the fastener. Because of the nature of impact loading of an impact wrench compared to constant loading, such as a drill, an impact wrench can deliver higher torque to the fastener than a constant drive fastener driver.
Ordinarily, a socket is engaged with a polygonally-shaped mating portion of the anvil of an impact wrench, usually a square-shaped portion, and the socket is, in turn, coupled to a polygonally-shaped portion of a fastener, often having mating hex geometry. The socket commonly has a polygonal recess for receiving the polygonal portion of the fastener, thus resulting in a selectively secured mechanical connection. This connection or engagement of the socket to the fastener often affords some looseness allowing for ease of repeated and intended engagement and disengagement of the components because of tolerance clearances or gaps between the components, wherein the gaps can vary in dimension, possibly as a result of manufacturing variation, and affect the timing and/or a spring effect commonly associated with the transfer of energy from the socket to the fastener. Additionally, there is often also a spring effect between the ordinary square-shaped socket and anvil mating connection. Therefore, it is desirable to increase the amount of torque applied by the socket to overcome spring effect, to maximize energy transfer, to increase net effect, and to improve performance of the impact wrench.
An aspect of the present disclosure includes an impact wrench comprising: a housing, configured to house a motor; a hammer, configured to be driven by the motor; an anvil configured to periodically engage the hammer as it is driven; and a socket having an interface configured to be removably coupled to a corresponding interface of the anvil, wherein the socket is further configured to engage a fastener; and wherein the anvil and socket are tuned and configured so that their combined stiffness, when removably coupled together including the interface between the two, is optimized so as to be between 1.15 and 1.45 times the stiffness of the fastener upon which the impact wrench is being used.
Another aspect of the present disclosure includes an impact wrench comprising: a housing, configured to house a motor and a hammer driven by the motor; an anvil configured to periodically engage the hammer as it is driven; and a socket removably coupled to the anvil, wherein the socket is further configured to engage a fastener; and wherein the anvil and socket are tuned and configured so that their combined inertia, when removably coupled together, is equal to the inertia of the hammer, thereby facilitating a hammer velocity of zero when the socket exerts peak force upon the fastener during tightening.
Still another aspect of the present disclosure includes an impact wrench comprising: a housing; a motor within the housing; a hammer driven by the motor; an anvil configured to engage the hammer; and a socket removably coupled to the anvil, wherein the socket is further configured to engage a fastener; and wherein the anvil and socket are dynamically tuned and configured so that the ratio of the inertia of the combined socket and anvil components and the inertia of the hammer has a specific relationship with the ratio of the anvil/socket combination stiffness and hex stiffness to achieve maximum output at a minimum total weight.
Yet another aspect of the present disclosure includes a method of dynamically tuning the drive components of an impact wrench, the method comprising: modifying the interface between an anvil and a socket so that the combined stiffness of the anvil and socket when coupled together is in the region of 4/3 the stiffness of the hex fastener on which the impact wrench is being used.
A further aspect of the present disclosure includes a method of dynamically tuning the drive components of an impact wrench, the method comprising: modifying the weight distribution of an anvil and a socket so that their combined inertia, when removably coupled together, is equal to the inertia of a hammer of the impact wrench, thereby facilitating a hammer velocity of zero when the socket exerts peak force upon the fastener during tightening.
Still a further aspect of the present disclosure includes a method of dynamically tuning the drive components of an impact wrench, the method comprising: equating the drive components of the impact wrench with springs and masses in a double oscillator model so that a hex fastener is equated with a first spring force, a socket is equated with a first inertial mass, an anvil is equated with a second spring force, and a hammer is equated with a second inertial mass; and tuning the anvil and socket so that the ratio of the inertia of the combined socket and anvil components and the inertia of the hammer has a specific relationship with the ratio of the anvil/socket combination stiffness and hex stiffness to achieve maximum output at a minimum total weight.
The foregoing and other features, advantages, and construction of the present disclosure will be more readily apparent and fully appreciated from the following more detailed description of the particular embodiments, taken in conjunction with the accompanying drawings.
Some of the embodiments will be described in detail, with reference to the following figures, wherein like designations denote like members:
Referring now specifically to the drawings, an example of a prior art impact wrench and a common a socket, is illustrated and shown generally in
A common socket 1010 ordinarily has a longitudinal axis 1028 that defines the rotational axis of the socket 1010 when it is secured to the socket engagement portion 1014 of the anvil 1022 of the impact wrench 1012. The socket 1010 also includes a body 1030 that extends along the axis 1028 from a first longitudinal end 1032 to an opposite second longitudinal end 1034. An input recess 1038, which is sized to receive and mate with the socket engagement portion 1014 of the anvil 1022 of the impact wrench 1012, is defined at the first longitudinal end 1032 of the socket body 1030. Typically, the recess 1038 is square-shaped to match the standard square-shaped cross-section (see
The socket 1010 normally includes an output recess 1040 that is defined at the opposite second longitudinal end 1034 of the body 1030. The output recess 1040 is sized to receive a head of a fastener. Typically, the recess 1040 is hexagonal (see
As is well known by one of ordinary skill in the art, a typical impact wrench 1012 is designed to receive a standard socket 1010 and designed to deliver high torque output with the exertion of a minimal amount of force by the user. As shown in
Those of ordinary skill in the art appreciate that there are many known hammer 1020 designs, and also recognize that is important that the hammer 1020 is configured to spin relatively freely, impact the anvil 1022, and then spin relatively freely again after impact. In some common impact wrench 1012 designs, the hammer 1020 drives the anvil 1022 once per revolution. However, there are other impact wrench 1012 designs where the hammer 1020 drives the anvil 1022 twice per revolution. The partial cut-away view of the impact wrench 1012 depicted in
The output torque of an impact wrench, such as impact wrench 1012, can be difficult to measure, since the impact by the hammer 1020 on the anvil 1022 is a short impact force. In other words, the impact wrench 1012 delivers a fixed amount of energy with each impact by the hammer 1020, rather than a fixed torque. Therefore, the actual output torque of the impact wrench 1012 changes depending upon the operation. An anvil, such as anvil 1022 or 3022 is designed to be selectively secured to a socket, such as socket 1010. This engagement or connection of the anvil, such as anvil 1022, 3022, to the socket, such as socket 1010, results in a spring effect when in operation. This spring effect stores energy and releases energy. Additionally, there is a spring effect between the socket 1010 and the fastener 1 to which it is engaged. Again, this spring effect stores energy and releases energy.
It may be beneficial to model the spring effects associated with tightening fasteners using an impact wrench. As is known to one of ordinary skill in the art, the combination of two masses (m1 and m2) and two springs (k1 and k2) is often referred to as a double oscillator mechanical system. In this system, the springs (k1 and k2) are designed to store and transmit potential energy. The masses (m1 and m2) are used to store and transmit kinetic energy. The drive system or drive components and mechanisms of common impact wrenches can typically be broken down into common fundamental elements. Ordinarily, the drive system is composed of a motor, a hammer, an anvil, a socket, and a joint (or fastener component that is to be driven). The motor can be directly or indirectly coupled to a hammer. The hammer often engages an anvil having mating jaws spaced apart from the center of rotation. The anvil is coupled to a socket with a mating geometric shape, usually a square, and the socket is usually coupled to the nut of the joint with mating hex geometry. As depicted in
A common impact wrench drive system employing a standard swinging weight or Maurer mechanism is particularly depicted and modelled in
For purposes of modelling, the common square drive anvil inertia is extremely low relative to the other components and is treated purely as a torsional spring. The compliance of the drive connection between the socket and the anvil is lumped into the total stiffness of the rest of the anvil and, for purposes of further modelling, will be assumed to be included in the term “anvil stiffness” and will be discussed later. The socket, such as socket 1010 is of relatively high stiffness but relatively large in inertia and is therefore treated as a pure inertia. For the sake of mathematical modeling the joint (or hex fastener 1) is assumed to be in the “locked” condition, i.e. unable to be moved further, allowing the hex interface to be modeled as a very stiff spring. It is the point at which the tool cannot move the hex any further that will characterize the “power” of the system. This is true in practice as well. A weak tool normally reaches a locked hex in a relatively short angle and the installed torque is low, whereas a strong tool normally reaches locked hex in a larger angle and achieves a higher installed torque on the same bolt.
The double oscillator system can be tuned to efficiently and effectively transfer energy from the impact device or hammer (modelled as m2) through the anvil-socket connection (modelled as k2), the socket (modelled as m1) and socket-fastener connection (modelled as k1) and into the joint fastener 1. Proper tuning can help ensure most of the energy delivered by the impact wrench hammer m2 is transferred through the anvil-socket connection spring k2 and into the socket m1. During use, the rate of deceleration of the inertial mass of the socket m1 is very high since spring k1 is stiff. Since deceleration is high the torque exerted on the fastener is high.
One way to tune the drive components of an impact wrench is to increase the inertial mass of the socket; to create a power socket. This can be done, inter alia, by providing the socket with an inertial feature, such as for example an annular ring located a radial distance away from the central axis of the socket. As depicted in
In reference to the disclosed tuned power socket embodiments, as illustrated in
When impact wrench drive system tuning is focused primarily on the socket, the tuning process operates under the notion that there is an optimal socket inertia for a given combination of mechanism inertia and joint and anvil stiffness. As such, the elements of the double oscillator system are predetermined. The rotary hammer inside the impact wrench m2 and springs k1 and k2 are assumed to have defined values. For tuning the system with the focus primarily on the socket, the only value which needs to be determined is the inertia member m1 2036 of the socket 2010, in order to achieve socket-optimized inertia. A common impact wrench, depending upon the drive size (i.e. ½″, ¾″, 1″), has a different optimal inertia for each drive size. The spring rate k2 and the rotary hammer inside the impact wrench m2 are substantially the same for all competitive tools incorporating common drive mechanisms, such as, for example, those impact wrench drive systems depicted and modelled in
To fully, and even optimally, tune an impact wrench drive system, focus may also be placed on the anvil and socket combination, two important components of an impact wrench drive system, and tuning methodology may consider optimizing the characteristics of each of the impact wrench drive system components that function together, to not only to have a stronger interconnection between the parts but to also perform at a higher level without introducing additional power input. Such optimal impact wrench tuning methodology introduces the concept of dynamic manipulation of both the socket inertia and the anvil-socket stiffness, in order to minimize the socket inertia for maximum output, thereby minimizing total tool weight and size. Dynamic impact wrench tuning, therefore, contemplates the ratio of the inertia of the combined socket and anvil components, as well as the inertia of the impacting mechanism, and considers how drive system performance has a specific relationship with the ratio of the anvil/socket combination stiffness and hex stiffness to achieve maximum output at the minimum total weight. The theory behind tuning the power socket, and in particular the methodology associated with determining the optimal component inertia of the socket still applies. The difference is the introduction of an additional independent variable.
When dynamically tuning impact wrench drive components, focus may be placed on the behavior of the various drive system elements, when in contact as a result of the collision of the hammer with the anvil from the moment of first contact until the moment the energy has reached and transferred to the bolt hex. At the beginning of this energy transfer period the hammer inertia has some initial velocity which represents all the kinetic energy that any particular impact can possibly have. When initial contact between the anvil and hammer jaws occurs, there is ordinarily a measurable amount of rotational clearance between the engaged components that must be consumed before any energy can transfer. There may be rotational play between anvil and socket, particularly if that connection is facilitated by the common square-shaped geometry. There is also rotational play between the internal hex of the socket and the external hex of the nut. Depending on how one chooses to consider the rotational clearance typically existent between impact wrench drive system components, there are two primary tuning models that may be implemented to fully, and even optimally, tune the impact wrench drive system. The optimal cases for each of the tuning models serve to provide upper and lower bounds for dynamically tuned impact wrench drive system performance.
For the purposes of this model and related discussion, an assumption is made that rotational play or clearance gaps between impact wrench drive system components has no significant effect on the behavior of the drive system and is assumed to be completely consumed. As the impact wrench drive system components wind up, all of the mechanical elements having varying amounts of inertia and stiffness contribute to a relatively complex oscillatory behavior. Energy is transferred from each spinning inertia to each series spring element, and kinetic energy converts to potential energy and back again in what may seem somewhat chaotic in the span of milliseconds. Tuning methodology focused primarily on modifying a socket to create a tuned power socket has taught us that choosing the inertia of the socket to be substantially higher than that of currently available standard sockets can enhance the transfer and concentration of the energy into the joint without increasing the energy put into the system. Understanding the relationships between these parts and the effects of their inertias and associated stiffness when interacting with each other, and the delivery of energy through the system, is critical to dynamically optimizing the impact wrench system to deliver as much energy to the fastener joint as possible.
The dynamic tuning and optimization process for socket and anvil inertia and stiffness of the various component connections of the impact wrench/fastener joint system begins with a calculation of the system modeled as lumped masses and springs where there is no rotational play or clearance gaps in between the components and the components are connected rigidly when they first come into contact. For the energy transfer time period in question, this assumption is reasonable and helpful to simplify the motion formulas. A typical schematic diagram for modelling a standard air driven impact wrench is shown in the diagrams depicted in
As discussed previously, a typical square-shaped anvil/socket mating connection has relatively low inertia, and the compliance of the anvil/socket connection is lumped into a total “anvil stiffness.” In an ideal case where the designer has complete control over all elements of the system, including the hex, there is a closed form solution to the positions, velocities and accelerations of the spring-mass oscillators shown in
x
2
=C
1α21 sin(ω1t+φ1)+C2α22 sin(ω2t+φ2)
x
1
=C
1α11 sin(ω1t+φ1)+C2α12 sin(ω2t+φ2) Equations 1
The initial conditions of the impact wrench drive system are given as:
For this set of initial conditions, the constants “a” and “C” are as follows:
The phase angels Φ and zero and the a's describe modal shapes:
The following assignment may be made:
α11=α12=1
Which, then reduces the “C” and “a” constants to:
Where the natural frequencies ω1 and ω2 are given by:
Equations 1 to 6 describe the motion of the mass under some initial conditions and any set of spring constants and inertias. In the ideal case where the designer has control of all inertias and stiffnesses, the specific values of those quantities can be determined by applying some dynamic energy accounting conditions throughout the impact cycle. Maximum deflection of spring k1, or the peak energy in the hex, would occur when all other components had completely given up their energy at the precise time when k1 reached its peak energy. This means that the hammer and socket have no kinetic energy and hence zero velocity and the anvil, spring k2, has no potential energy and hence, no deflection. Again, this is the ideal case.
To find the optimal torque in spring k1, the following conditions may be applied to Equations 1-6:
At some subsequent time t=A
For this set of conditions to be true:
m
1=¾m2k1=¾k2
The above results describe the optimal inertia and the optimal stiffness of the components that reside between the hammers and the hex under ideal conditions. Hence, when dynamically tuning an impact wrench drive system, in a perfect world, the inertia of the socket/anvil combination must be ¾ of the combined inertia of the hammer components of an impact mechanism at the same time the stiffness of the nut hex must be ¾ of the stiffness of the anvil for maximum output and minimum total weight.
It may be helpful to visually demonstrate the difference between a standard impact wrench drive system having a square-shaped anvil/socket connection, such as the impact wrench drive systems depicted and modelled in
With regard to the dynamically tuned and optimized impact wrench drive system,
While the inertia of the socket can readily be increased or decreased through introduction of part-geometry changes, such changes may result in unwanted adverse effects on the overall weight of the tool and the ability to access tight spaces where bolts or other fasteners might be located. Achieving the optimum stiffness is much more challenging than the ideal case for at least two reasons: 1) there are many nut sizes in existence on which the impact wrench will likely be used, which presents the possibility for a wide range of stiffness ratios to a given tool—a decision must, therefore, be made regarding a hex size for which to optimize; and 2) the anvil stiffness, which includes the stiffness of the interface between the anvil and the socket can be quite low, as in the currently available common square-shaped interface impact wrenches, such as those with drive systems depicted and modelled in
1/KTotal=1/Ksquare+1/Kanvil
K
Total=1/(1/Ksquare+1/Kanvil) Equations 7
So, when the lab-measured data for a common Maurer mechanism impact wrench with a standard square-shaped interface (See
K
Total=1/(1/274,000+1/55,000)=46,000
In order to achieve an optimal stiffness ratio, the total anvil stiffness (including the interface with the socket) needs to be 4/3*K1. With regard to a 15/16″ hex fastener, as set forth in the data listed in Table 1 of
1/Kanvil=1/KTotal−1/Kspline
K
anvil=1/(1/KTotal−1/Kspline)
K
anvil=1/(1/446,700−1/1800K)
K
anvil=approx. 594,000 in-lb/rad
Obviously this stiffness is a significant increase over the standard anvil. However, a common cordless impact mechanism, often referred to as a “Ball and Cam” type mechanism, lends itself well to the geometry changes required to meet this requirement. The jaws of the corresponding hammer are spaced relatively far apart, which allows the anvil diameter to increase to not only better support the jaws, but a larger anvil diameter also increases the associated anvil inertia, thereby tuning the device and meeting the optimal inertia requirement. One such tuned anvil 5022 embodiment is depicted in various perspective views in
A tuned anvil 5022 is depicted, in
Dynamic impact wrench drive system tuning involves a determination, based on mathematical modelling as assisted by empirical data, of optimum trade-offs between inertia and stiffness. As depicted in
m
1=¾m2k1=¾k2
The Inertia Ratio vs. Stiffness Ratio plot can be very insightful for tuning purposes, especially when utilized in conjunction with empirical data pertaining to impact wrench drive systems. For example, as depicted in
When tuning impact wrench drive system performance through employment of spring-mass oscillation modelling, the Inertia Ratio vs. Stiffness Ratio plot can be used to determine the optimal inertia for ANY stiffness ratio that is achieved. There are performance advantages associated with moving the stiffness ratio as close to 1.33 as possible, so that the inertia can be as low as possible and still perform at the highest level. The stiffness of the interface between the socket and the anvil will determine the extent to which the required inertia can be split between the socket and the anvil. In the case of a square drive connection, both the model and empirical data have demonstrated that the connection is not stiff enough to treat the anvil and socket as a single mass and, therefore, the substantial part of the required inertia may be contained in the socket, such as in the design of the tuned Power Socket. With a stiffer connection, such as with the implementation of a spline drive connection, the inertia may be divided in any convenient manner between the two components thereby reducing the potential for reduced access due to the added material in the socket.
As set forth above, the ideal case is interesting, but it is the exception and not the rule. There are various reasons why the impact dynamics do not operate with all parameters at their optimum. Firstly, a user is likely to use an impact wrench on a wide range of hex sizes. Each hex size will exhibit a wide variety of stiffness behaviors. Large hexes will appear very stiff while small hexes will appear relatively soft. Estimating what stiffness to expect for a given hex size is a matter of experimentation and empirical testing. In the end, however, there will likely end up being hex sizes for which the tool is not fully optimized. The shape of the optimality curve, however, allows the designer to optimize at a relatively large hex size resulting in a very close to optimal condition for hex sizes below the hex for which the tool was optimized. Secondly, there are physical limitations to the anvil for each mechanism that may preclude the achievement of high levels of stiffness. For example, long and thin rods can typically be made stiffer by becoming shorter and/or larger in diameter. In the case of the anvil depicted in
The spring-mass oscillation model includes assumptions requiring the hammer, anvil, socket and hex nut fastener to be in contact for the duration of the impact event. Even if the forces arise during the simulation, the math of the model does not contemplate the separation of the component elements. Test data has revealed that this is actually a relatively rare case in actual practice, but certainly a possible and potentially bounding case.
There are several other possible combinations of state of contact that can affect the optimization of the inertia and stiffness parameters of an impact wrench drive system under a designer's control. The ability of the components to disengage from each other is commonly afforded by the loose fits between the anvil and the socket as well as between the socket and the hex nut. The loose fits are, to some extent, required in order to allow for manufacturing variation and ease of repeated and intended assembly and disassembly during normal use. These loose clearances or gaps between the impact wrench drive system components are critical to consider with regard to the energy transfer abilities of the drive system due to their more common presence at the time when hammer contact with the anvil is initiated.
The clearances between the impact wrench drive system components have an important role in the timing of the energy transfers that occur between the parts. In this momentum model analysis, it is helpful to think of the impact event, not as an instantaneous or discontinuous change in state, but rather a motion defined by rapidly changing accelerations that depend on what is in contact at that time. In the spring-mass oscillation model, the energy that each component contains during the impact event was described. In a zero-clearance scenario, as the spring-mass oscillation model described, there can be energy harbored in the various components at the time when it is desired for all of the energy to arrive at the interface between the socket and the nut. The harbored energy arrives late (if at all) and is unable to do any valuable work on the nut. In a clearance gap, or momentum model, scenario where there is an angle through which the socket (or anvil) is required to cross before any contact occurs, there can be substantially more time for the energy transfer between the bodies to occur BEFORE the nut/socket interface reaches its peak torque. A depiction of how the energy contained in the hammer, anvil, socket and hex at any given time for a given set of design and initial condition parameters is depicted in
The state of the clearance or gaps between the driving interfaces of the impact wrench drive system components is not currently controlled and is nearly random. However, the effect of the clearances on the optimization of the system is important. When there is time for full transfer of energy to take place, the optimization actually simplifies greatly especially if that time can be assured. From the spring-mass oscillator model, it has been established that the anvil stiffness has a distinct effect on the timing of the energy. This is a powerful parameter to use in order to improve how quickly and completely energy is transferred in the case of very low or zero hex clearance.
A plot of simulated torque output as the clearance in the hex, called “hex gap” increases is depicted in
The presence of clearance in the hex also effects the optimization of inertia. Whether the anvil stiffness is able to be increased or not, if separation between the anvil and the hammer is the predominant type of impact event, then the theories of collision and momentum will apply. Recalling the energy accounting and timing discussion of harbored anvil energy at low anvil stiffness, it is desirable to ensure that the hammer does not harbor energy (kinetic, in this case) when the torque in the hex reaches its peak. Otherwise, the energy is considered late and fails to contribute to the work done on the nut. Therefore, for optimal performance, the velocity of the hammer must be zero when the peak of the hex is reached. Assuming that other conditions are favorable for a disengagement scenario, such as a stiff anvil and/or an adequate hex clearance state, the proper inertia of the anvil/socket combination is equal to the inertia of the hammer. Consider the diagrams depicted in
As depicted in
The ratio of the masses dictates the ratio of the changes in velocities. Additionally, the continuing forward velocity of the striped ball and the spring toward which the white ball heads makes it likely that there will be additional contact between the balls before the striped ball's velocity becomes negative and heads in the opposite direction from which it approached. This bouncing behavior is highly inefficient and undesirable operation.
If the white ball is significantly smaller than the striped ball then the striped ball will have a continuing (positive) velocity in the original approach direction, as in
The only way to get complete momentum and energy transfer is when the striped ball has zero velocity after the impact. Knowing that Vwi and Vsf are both zero and that Vsi is NOT zero, the only way for this to be true is for ms and mw to be equal therefore making Vwf=Vsi.
Therefore, for cases where the state of hex clearance is adequately large or the anvil is relatively stiff, the optimal socket/anvil inertia is equal to the hammer inertia, as depicted in
Since the hex clearance state at any impact event is relatively random, there will be conditions that will vary between a zero gap condition and an adequate gap condition. Prediction of performance will then have an upper bound defined by the momentum-based model and a lower bound defined by the spring-mass oscillation model, as depicted graphically in
An impact wrench having dynamically tuned drive components may be capable of generating higher torque outputs, without increasing the weight, size or cost of the tool. The tuned drive components are optimized for inertial performance and stiffness, and are capable of transmitting energy more effectively and efficiently than standard impact wrench and socket designs. As such, a dynamically tuned impact wrench may solve the problem of achieving both high impact torques while operating a maximum motor operating points, and may also prevent erratic operation while being operated at low mechanism speeds. The tuned drive components permit successful performance in both modes of operation (max motor and low speed), while incorporating lighter weight componentry having smaller size requirements. Another advantage obtained from utilizing an impact wrench having dynamically tuned drive components is a substantially advanced combination of extreme impact power and untethered portability. For example, extreme impact power has been obtained before, but has always been limited to pneumatic powered applications, which require and air hose to be connected to the tool and thusly restricting tool mobility. Dynamic tuning of the drive components facilitates the integration of reduction gearing in the drive train and permits the motor to more effectively run at high speed. Moreover, a tuned wrench obtains the advantage of increased power to weight ratio, since standard parts can be reduced in size, while still maintaining qualities of high performance and durability.
When dynamically tuned, the socket and anvil are still separate components but are connected by an extremely stiff connection, such as a spline. The stiffness of the connection between the two drive components provides the following advantages over present solutions: 1) it causes the connected components to substantially behave as a single component such that the inertia of the anvil can simply be added to the socket inertia when determining optimal inertia. This means all the inertia required for optimal performance does not have to exist on the socket itself, but can be “hidden” further back inside the tool out of the immediate region of the fastener; and 2) a limiting factor in the overall stiffness of the anvil-socket combination has typically been the square shaped connection between the socket and the anvil. Increasing the socket/anvil connection stiffness allows the overall stiffness to be increased. As mentioned previously, increasing this stiffness reduces the inertia required to reach optimal performance. Tuning methodology, implemented through execution of at least one of two primary models (the spring-mass oscillator model and the momentum model) has rendered optimized performance characteristics with bounded ideal cases allowing for introduction and comparison or empirical test data, thereby facilitating part design optimized for multi-varied tool operation differences, such as looseness or clearance gaps between coupled components, as well as optimal structure changes in view of the balance between inertia ratios and stiffness ratios. For example, dynamic tuning reveals that the total stiffness of the anvil-socket combination including the interface between the two is in the region of 4/3 of the stiffness of the hex on which the tool is being used. Otherwise the inertia ratio for optimal performance at the minimum weight is a prescribed value in relation to the stiffness ratio.
As depicted in
The tuned anvil 6022 may be mated to a correspondingly tuned socket 6010, as depicted in exploded view in part of
Similar modeling alteration is depicted in
As discussed, there are several advantages obtained from dynamically tuning the drive components of an impact wrench. For example, one such advantage pertains to desirous changes in the external dimensions of tuned components, as depicted in
The engagement structure between the dynamically tuned socket and anvil has been primarily described and depicted as an involute spline with teeth that standard cutting tools in the industry can manufacture. A spline engagement is, therefore, desirable from the standpoint of both manufacturability and strength. However, there are alternatives that can also meet (or come close to meeting) the stiffness, inertial, and durability necessities pertinent to dynamically tuned impact wrench drive components. For example,
While this disclosure has been described in conjunction with the specific embodiments outlined above, it is evident that many alternatives, modifications and variations will be apparent to those skilled in the art. Accordingly, the preferred embodiments of the present disclosure as set forth above are intended to be illustrative, not limiting. Various changes may be made without departing from the spirit and scope of the present disclosure, as required by the following claims. The claims provide the scope of the coverage of the present disclosure and should not be limited to the specific examples provided herein.
This application claims the benefit of priority and is a continuation-in-part of U.S. patent application Ser. No. 13/080,030 entitled ROTARY IMPACT DEVICE and filed on Apr. 5, 2011. This application also claims the benefit of priority and is a continuation-in-part of U.S. patent application Ser. No. 14/169,945 entitled POWER SOCKET FOR AN IMPACT TOOL and filed on Jan. 31, 2014. In addition, this application claims the benefit of priority and is a continuation-in-part of U.S. patent application Ser. No. 14/169,999 entitled ONE-PIECE POWER SOCKET FOR AN IMPACT TOOL and filed on Jan. 31, 2014.
Number | Date | Country | |
---|---|---|---|
Parent | 13080030 | Apr 2011 | US |
Child | 15290957 | US | |
Parent | 14169945 | Jan 2014 | US |
Child | 13080030 | US | |
Parent | 14169999 | Jan 2014 | US |
Child | 14169945 | US |