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Design considerations for the electric servo-motor driven 30 ton double knuckle press for precision forming
Int. J. Mach. Tools Manufact. Printed in Great Britain
DESIGN MOTOR
Vol. 33, No. 2, pp. 209-222, 1993.
0890-6955/9356.00 + .00 (~ 1993 Pergamon Press Ltd
CONSIDERATIONS FOR THE ELECTRIC SERVOD R I V E N 30 T O N D O U B L E K N U C K L E P R E S S F O R PRECISION FORMING S. YOSSIFONt and R. SHIveuPa$ (Received in final form 8 January 1992)
Abstract--This paper is the last part of a three part series on the design, analysis and construction of a flexible electric servo-motor controlled mechanical press for precision forming. In Part I, characteristics of the available mechanical drives for presses were compared and a double knuckle drive selected for this press application. This drive was shown to have the desirable characteristics of a large mechanical advantage and a constant ram velocity over the working stroke. Part II presented an enumeration search procedure for optimization of the link design for maximizing the mechanical advantage at the working stroke, with the constraint of a corLstant working velocity. This procedure was applied to the double knuckle linkage drive and the optimized dimensions selected. This paper presents the application of the above procedures to the development of a 30 ton double action press. Focus of this paper is on the scaling of the optimized design, force and torque analysis and the selection of press specifications. INTRODUCTION
MECHANICALpresses have been widely employed for the sheet and bulk metal-forming processes. However, due to their high velocities in the working zone, limited working strokes and poor controllability they have lost ground to hydraulic presses for cold forging and deep drawing applications. Recent attempts in the mechanical press design have focused on using a linkage drive mechanism to increase the approaching and return speeds, but to slow down the ram in the working zone. Another design objective is to maintain the specified load over a relatively long working stroke, while simultaneously reducing the crank torque and the frame strength requirements to achieve reductions in the weight, size and cost of the machine. Many different types of linkage drives proposed to achieve these objectives, have been studied [1]. Recently developed press drives include the double-toggle linkage, four-bar linkage, link type, drag crank linkage and the toggle lever presses. In all these drives, the idea was to satisfy the load-stroke characteristics for the specific metal-forming process. A 2.5 ton multiaction press driven by an a.c. servo-motor for precise position and velocity control was developed earlier at the Engineering Research Center (ERC) [2]. In this press, the angular displacement of the servo-motor was converted to a linear displacement by a worm-gear ball-screw jack. Recently, this press was modified for sheet metal-forming [3]. In parallel, two types of double-toggle linkage mechanisms were studied and proposed by Nye and Shivpuri [4]. In the first design, the motor and the screw were', held stationary as the linkage traveled through its stroke (type A in Fig. 1). The second arrangement eliminated the connector link by allowing the motor and the screw to pivot as the linkage traveled (type B in Fig. 1). A double-toggle rotary press with crankshaft (type C in Fig. 1) was proposed by Yossifon et al. [5]. The rotary drive (crankshaft) is capable of transmitting a higher torque (with the same motor) than the linear drive owing to its higher efficiency. Use of an a.c. servo-motor on the linkage press provides for the independent control of ram position and velocity. It can easily be interfaced with C N C machine controls because of its extensive use in the machine tool drives. The servo-motor is also capable of transmitting near constant torque over its speed range [6]. Control of the ram speed,
?Materials Division, NRCN, POB 9001 Beer-Sheva, Israel. SDepartment of Industrial and Systems Engineering, The Ohio State University, Columbus, OH 43210, U.S.A. 209
210
S. YOSSlFON and R. SHIVPURI
+Y +X
TypeA
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FIG. 1. Schematic of type A, type B and type C design linkages.
independent of the other characteristics, provides more flexibility for the press user. The speed profile can be predetermined and programmed for each process to match the load-stroke characteristics of the press. Using a rotary transmission (crankshaft) driven by an a.c. servo-motor provides two load-stroke curves for each direction (clockwise or counterclockwise). In addition, the stroke length, bottom dead center (BDC) and top dead center (TDC) can be predetermined and programmed by rotating the crankshaft for only a portion of the cycle, and thereafter return back to the home position in the opposite direction. In the previous study [7], kinematic simulation programs were used for comparison between ERC rotary drive and four well-known rotary linkage drive mechanisms selected from the literature. The objective functions were the mechanical advantage and linkage drive size for the same total ram stroke. It was obtained from the kinematic
Design Considerations for the Double Knuckle Press
211
simulation that the mechanical advantage of the ERC rotary press is near constant along the major ram stroke. Therefore, the rotary double-toggle mechanism was chosen for prototype construction and optimization. For the verification of the kinematic simulation and design concepts, a wooden prototype of the press and a control unit were designed and built at the Center. The selected rotary double knuckle joint mechanism was then optimized based on design criteria and feasibility of the design [8]. Two objective functions (design criteria) were established: (a) the mechanical advantage of the mechanism; and (b) the size ratio (linkage dimensions non-dimensionalized with respect to the total stroke of the ram). In this study, a cyclic coordinate search method was implemented [9]. Several enumeration searches for the best arranged mechanism were completed. Based on these searches, an optimized rotary press design for a 5 in ram stroke was selected (Fig. 2). The goal of the current study (Part III) is to design an a.c. servo-motor driven 30 ton optimized double knuckle mechanical press. This press will be capable of controlled bulk forming experiments using model materials and sheet forming experiments such as thin sheet drawing. This press design procedure includes: rescaling and repositioning of the optimized rotary linkage drive for final comparison and selection; and force analysis of the optimized rotary linkage drive in order to evaluate the bearing forces and the required input links torque. From the analysis, preliminary construction design specifications are proposed. ANALYSIS OF THE LINKAGE DRIVE
Three different programs were used to study and compare the motion characteristics of the selected linkage drives. The first program, KINLYN [10], is capable of analyzing 2-D linkages for position, velocity and acceleration/deceleration. The second program, CATIA [11], is a general purpose CAD/CAM software for graphical display and analysis. The third program was a F O R T R A N code that read in the data file (ram stroke vs crank angle) and performed a finite difference incremental technique to find the derivative of the ram stroke vs input angle (in radians) curve for a rotary drive
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Node 1 (10.5607,-4.2968) Dimensions of the Optimized Press for Five Inch Stroke L2 L3 L4 L5 L6 L7
= 3.55917" = 8.82199" = 7.82290" = 7.52993" = 5.39344" = 8.55578"
Node 4
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FIG. 2. Schematic of the optimized ERC press for a 5 in ram stroke.
Node 7
212
S. YOSS1FON and R. SHIVPURI
and the derivative of the input stroke vs output stroke for a linear drive. The computer program calculated the necessary values, and generated a plot of mechanical advantage (MA) vs dimensionless stroke (h/H) curve for the linkage mechanism. Details on the use of these programs for kinematic analysis are presented in Parts I and II [7, 8] of this series. In our previous study [8], the optimal arrangement and the dimensions were obtained after several enumeration search steps for the ERC press (type C in Fig. 1) for a 5 in total stroke (see Fig. 2). The mechanical advantage improved from 2.5 to 4.0 due to the optimization of the linkage, this optimization also resulting in the size of the linkage mechanism being reduced for the same ram stroke. Details of this procedure were presented earlier [8]. Since the dimensions of the press are influenced by the arrangement of the links, it was decided to investigate the best location for the crankshaft center (CC) while keeping the link dimensions the same. In Fig. 2, the original location of the CC is at node 1, which is 17.2526 in to the left and 3.5254 in below the position of node 6 (the highest vertical point). As a first iteration, the CC was repositioned to a new location as in Fig. 3(a) (13.7 in to the left and 8.866 in below node 6) keeping the link dimensions and pivots the same. This new location substantially saves the space requirements. The output data file for the new design is plotted in Fig. 3(b). The total ram stroke obtained is about 3.5 in, instead of 5 in before repositioning. Therefore, this repositioning did not improve the motion performance of the optimized rotary linkage drive (type C). Further investigations of the location limits of CC were conducted using the linear arrangement (type A, Fig. 1), since it has similar characteristics to the rotary drive (type C, Fig. 1). Press A design was run on the CATIA program and the link positions along the entire ram stroke were plotted. Two positions of the angle of the linear input were selected as extremes of CC movement without changing the drive characteristics: 18.68 ° and 70° relative to the horizontal, Figs 4 and 5. Increasing the input angle significantly reduces the mechanism width which in turn reduces the overall size of the mechanism. However, the mechanical advantage of the "modified" linear press is near constant along the working stroke and too low, about 1.75 (Fig. 6).
FIG. 3(a). Links positions along the entire ram stroke of the repositioned optimized ERC rotary press.
Design Considerations for the Double Knuckle Press
213
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C R A N K ANGLE (deg) Fro. 3(b). Ram stroke vs crank angle of the repositioned ERC rotary press.
18 ° 41'
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TDC
BDC
FIG. 4. The links positions of press A along the entire ram stroke. PRELIMINARY DESIGN OF THE LINKAGE DRIVE PRESS
The primary objective of the current study is to design and build a 30 ton press with the optimized double knuckle joint mechanism, driven by an a.c. servo-motor for accurate control of the velocity and the position of the ram. The following initial parameters were established for the press: (i) (ii) (iii) (iv)
servo-motor output torque of 240 in-lb; maximuJrn motor velocity is 1200 rpm; the total[ stroke of the ram is H = 5 in; and the required ram force is 20 tons at 1.5 in (h/H = 0.3) above the BDC.
With these parameters, the dimensional analysis of the optimized press had to be completed to compare it to the characteristics of the linear arrangement (type A). The specifications of the a.c. servo-motor chosen to power the 30 ton double-toggle
214
S. YOSSIFON and R. SHWPUR1
FXG. 5. The links positions of the repositioned press A along the entire ram stroke.
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FIG. 6. The mechanical advantage of press A and the repositioned press A vs ram stroke.
press are given in Table 1 [6]. This was the largest capacity a.c. servo-motor available in the U.S. at the time of design. The torque-speed characteristics of the servo-motor presents near constant torque capability over its speed range. As a result of this unique feature and the linearity of the linkage drive in the working zone, the ram force is proportional to the rotary mechanical advantage ( M A * ) of the press for a constant torque input. For the
Design Considerations for the Double Knuckle Press
215
TABLE 1. SPECIFICATIONS OF THE SERVO-MOTOR K H X - 7 4 0 FOR THE 30
TONPRESS[6] Continuous stall torque Peak torque Rated power Rated speed Rated motor stall current Rotor inertia Acceleration at peak torque Maximum resolution Motor weight
256.7 lb-in 513.0 lb-in 4.56 hp 1300 rpm 20 A 210.5 oz in-2 15,061 rad s-1 32,768 steps rev-1 58.3 lb
continuous duty region, the average output torque is 240 lb-in and the peak torque is limited to 513 lb'-in, A program was written that rescaled the linkage drives to match the required parameters and plotted desired curves of force and velocity. The stroke of the desired optimal press was chosen to be 5 in, a typical stroke needed for cold impact extrusion and sheet metal drawing of medium size components. After rescaling the lengths, the program performed a single position analysis to obtain the angle vs stroke data. The linear mechanical advantage data (MA) was calculated through the incremental slope method [7]. Knowing the necessary force and the desired mechanical advantage value at a certain point from the BDC, the necessary output torque to drive the linkage was calculated. The ram force F is determined by: F = (MA*) (M)/H
(1)
where M is the,, input link torque and H is the total stroke (5 in). The angular velocity of the input link was found by using the angular velocity of the servo-motor and dividing it by the gear ratio of the gearbox connected between the servo-motor and the linkage drive crankshaft. The gear ratio is calculated easily by determining the ram force required at the beginning of the working stroke. To find the linear velocity of the ram, the chain rule was used: dh dt
dh dot d a dt "
(2)
The ram velocity is dh/dt, and the value of det/dt is simply the angular velocity of the input link. The values of dh/dcL were calculated using an incremental slope method as presented in Ref. [7]. The rescaling of the linear drives is calculated through the following procedure. The ram force (Fout) is calculated as: Fo,t = (MA) (Fin)
.
(3)
The input force at the crank (Fin) is constant and is obtained through the following equation [12]:
Fin-
2'rrMi. L
(4)
where Min is the servo-motor output torque (constant torque) and L is the screw lead (in t u r n - l ) . To find the ram velocity, the chain rule from equation (2) had to be used. The ram velocity is dh/dt and the value of doddt is the angular velocity of the servo-motor. The value dh/dot is as follows:
216
S. YOSSIFON and R. SHIVPURI
dh da
L MA "
(5)
By substituting equation (5) into equation (2), the ram velocity obtained was:
L( o)
Vou,=M~ ~ •
(6)
The decision was made from design considerations, that at 1.5 in stroke from the BDC (nominal point) the required ram force is 20 tons. Knowing the mechanical advantage value over the total ram stroke, the output force and speed were predicted (using a.c. servo-motor type KHX-740 [6]). The values of force and velocity of the linear arrangement (type A, Fig. 1) and the repositioned linear press were plotted. One can see that for type A, the ram force curve at the working stroke rises up sharply toward BDC, and the ram velocity drops down non-constantly over its total stroke (Figs 7 and 8). For the repositioned type A, the ram force and the ram velocity curves were more acceptable. These results indicate that raising the CC increases the mechanical advantage but decreases the region of constant force and torque, while lowering CC does the opposite. The curves generated for the ERC optimized press are shown in Fig. 9(a) and (b) for ram force and ram speed respectively. The ram force at the working stroke is near constant and decreases moderately in the approaching stroke (Fig. 9(a)). Also the ram speed is near constant in the working stroke ( - 0 . 8 in s -1) and increases sharply at the approaching stroke for a constant input speed of 1200 rpm (Fig. 9(b)). These characteristics provide for (a) a near constant ram force and velocity in the working region and (b) rapid approach speed for reduced cycle time. Additional attempts to modify press C by repositioning the CC while using the same link lengths and geometry showed that press C with the location of Fig. 2 is the best design from the mechanical advantage, force and velocity characteristics, and the overall linkage size points of view.
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FIG. 7. The maximum ram force of press A and the repositioned press A vs ram stroke.
Design Considerations for the Double Knuckle Press
217
4.2 Rep. Type A Type A 3.2
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RAM STROKE ( i n ) F]o. 8. The maximum ram velocity of press A and the repositioned press A vs ram stroke. FORCE ANALYSIS OF THE SELECTED LINKAGE DESIGN
After performing a kinematic analysis, optimization and rescaling of the doubletoggle rotary mechanism, a kinetostatic force analysis was done. This analysis took into account the inertial forces imposed on the links. The objective of the kinetostatic analysis is the de,termination of the bearing forces and the required input crank torque at the particular instant in the position being analyzed. Once analytical expressions are found for a single position, magnitudes of bearing forces and input torque for additional positions are easily determined by repeating the analysis for these additional positions. The matrix method of kinetostatic analysis was used to obtain the analytical expressions. This method considers all the inertial forces together, as compared to the superposition method wherein the effect of individual inertial forces are treated separately and then superimposed to determine their combined effect. One advantage of the matrix method is that the equations of motion are quickly derived. The most fundamental technique in force analysis is to break down a mechanism into several free bodies with compatible forces and torques applied to each free body. In each free body diagram of a link, the inertial forces are assumed to act at the center of mass and an inertial torque is added [13]. The mechanism is analyzed at each position by D'Alembert's principle: ~'~Fx + ( - M agx) = 0
(7a)
EFy + ( - ] ~ agy) = 0
(7b)
~,T + (-Igak) = O .
(7c)
In the above equations, ( - M agx) is known as inertial force component in the x direction, ( - M agy) is the inertial force in the y direction and ( - I g ak) is the inertial torque. The accelerations agx and agy a r e specified by subscript "g" denoting that these are accelerations of the centers of masses in the x or y directions. Based on the Cartesian coordinate system, the x, y components of forces, positions and accelerations are HTH 33:2-H
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Fro. 9. (a) R a m force curve vs ram stroke for the E R C optimized press. (b) Maximum ram velocity vs ram stroke for the E R C optimized press.
assumed positive in the positive x and y directions, respectively• The counterclockwise direction is assumed as the positive torque direction• For each link, one can write three equilibrium equations (equations (7a), (7b) and (7c)), and for each free node (node numbers 2, 3 and 5 in Fig. 2) two equilibrium force equations (for the x and y directions)• Thus, there are 24 equations and 24 unknowns, and this system of equations can be solved• Since the positions of the nodes are known from the kinematic analysis, the system of equations is linear in the unknowns and can be expressed in the matrix form:
[L][FB] = [FI]
(8)
where [L] is the square matrix of known linkage parameters, [Fa] is the column matrix of unknown joint forces and input torque, and [F~] is the column matrix of known external loads plus inertial forces and torques. The solutions were solved using a linear equation solver by implementation of the Gauss elimination method• From design considerations, the linkage strength was to be calculated for a maximum ram force of 50 tons (approximately the maximum possible
Design Considerations for the Double Knuckle Press
219
force generated by the linkage in its entire stroke). These equations were then solved assuming a con,;tant ram force of 50 tons throughout the working stroke, and the torque, which was to be supplied by the gearbox unit through the servo-motor, was checked. However, since the servo-motor integrated with a transmission unit of 205 gear ratio did not have the capacity to supply torque over its rated maximum torque capacity, which is limited to 41.0 ton-in (400 lb-in motor torque) by a special adjustable clutch (the continuous permissible crank torque is limited to 24.6 ton-in), it would be able to supply only that maximum torque and the ram force that would be available would drop below its 50 tons capacity. The factor by which the servo-motor's rated torque capacity fell short of the torque requirement imposed by the constant ram force of 50 tons was used to accordingly reduce all the forces. Thus, the working stroke was simulated as two load steps. Near the BDC, links 6 and 7 (Fig. 2) were nearly vertical and this allowed a large fraction of the full load to be transmitted through the links to the ground. However, as the links moved away from the BDC, the mechanical advantage of the mechanism dropped and this led to very high torque requirements on the servomotor. Now the., condition used to solve the equations was that of constant torque supply by the servo-motor. Hence, the working region was divided into two regions, one where the ram force would be a 50 tons constant (curve "a" in Fig. 10) and the other when the torque transmitted by the servo-motor and the gear unit was a 41 tonin constant (curve "b" in Fig. 10). Figure 11 shows the acting load in each individual link vs the crank angle along the working stroke. One can see that links 3, 5, 6 and 7 are subjected to compression forces along all of the working stroke, compared to link 4 which is subjected to tensile forces all over the working stroke. However, link 2 is subjected to tensile loads during a part of the stroke and compression loads during the rest of the stroke. The maximum link loads are presented in Table 2. These link loads will help determine the dimensions of the links and the joints of the drive linkage. In the detailed design of the joints, rigidity and wear considerations often override the strength requirements.
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300
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CRANK ANGLE, 0 (deg) FIG. 10. Plots of (a) input torque vs crank angle and (b) ram force vs crank angle, produced along the working stroke, when the maximum input torque is 41 ton-in and maximum ram force is 50 tons.
220
S. YOSSIFON and R. SHIVPURI 25 LINK 4
20
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~
-30
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260
280
300
320
340
360
CRANK ANGLE, e (deg)
FIG. l l . Link load vs crank angle in the working region for the linkage drive scheme shown in Fig. 2. TABLE 2. M A X I M U M LINK LOADS ALONG THE WORKING STROKE FOR MAXIMUM 5 0 TONS RAM STROKE
Link No.
2
Max. load (tons)
3
- 9 . 0 -14.7
4
5
6
7
24
-30
-50
-50
DESIGN SPECIFICATIONS FOR THE 30 TON PRESS
Based on the comparison and analysis [7] and the linkage optimization [8], a double action 30 ton capacity press is being built at the Engineering Research Center. The press uses a double knuckle joint linkage arrangement to drive the top punch with an a.c. servo-motor. The second action (clamping plate), is driven by four hydraulic pistons mounted on the top plate. The press will be used for precision bulk and sheet metalforming. The press consists of four circular guide posts, connected between the top plate and the base plate. The clamping plate travels with controlled velocity along the guide posts on bushings. The ram travels along the press center and is guided by bushing connected to the top plate. The angular displacement of the a.c. servo-motor is transferred to the ram through a reduction transmission and a double knuckle joint mechanism. The transmission system is mounted on the top plate. The conceptual drawing of the press construction is shown in Fig. 12. The focus of this paper being the mechanical drive, the details of the hydraulic drive for the clamping plate and the computer control system are not included. The planned press specifications and the performance are given in Table 3. CONCLUSIONS AND FUTURE WORK
In the current study it was proved theoretically that the orginal optimized rotary drive mechanism can be adopted for the 30 ton press design and construction. From kinematic and kinetostatic analysis, the motion characteristics and link loads were predicted for a KHX-740 servo-motor drive [6], which is available in the market. The designed press has a maximum actual ram force of 30 tons and maximum ram speed of 0.8 in s -1 at the working stroke.
Design Considerations for the Double Knuckle Press
221
Input from servomotor
or1 for linkage
G
loggle linkage
Rc,tar!
r bearing
)er platen Clan h guidance
I=ou
or lower punch
r platen
frame
FIG. 12. A schematic drawing of the 30 ton double knuckle joint press showing the mechanical drive components (side view). TABLE3.
PLANNED PRESS SPECIFICATIONS AND PERFORMANCE
Maximum ram force: Maximum clamping force: Maximum ram stroke: Ram fi~rce at 1.5 in stroke from BDC: Maximum clamping stroke: Maximum pistons working pressure:
30 tons 25 tons 5 in 20 tons 12 in 2000 psi
Maximum down stroke clamping velocity: Maximum up stroke clamping velocity: Maximum working ram velocity: Maximum ram velocity: Ram velocity profile:
0.55 in s 0.45 in s -~ 0.8 in s -~ 2.5 in s User specified/PC control
Maximum press height: Maximum press width: Maximum press depth: Minimam shut height: Daylight: Effective die set space:
~ 110 in ~ 50 in ~ 42 in 12 in 24 in 28 × 21 in
D e t a i l e d design o f th e 30 t o n press will be m a d e in t h e n e a r f u t u r e . T h e specifications o f t h e press witlh c o n c e p t u a l d r a w i n g s w e r e p r e s e n t e d earlier. T h e m o d i f i c a t i o n s o f a s e r v o - m o t o r d r i v e n m u l t i a c t i o n press f o r s h e e t m e t a l - f o r m i n g will b e i m p l e m e n t e d in t h e 30 t o n press design. M e a s u r e m e n t d e v i c e s such as a p u n c h l o a d d e v i c e , p r e s s u r e t r a n s d u c e r f o r c l a m p i n g f o r c e m e a s u r e m e n t an d a d e v i c e f o r m e a s u r i n g t h e p u n c h
222
S. YOSSIFONand R. SHIVPURI
d i s p l a c e m e n t will b e i n s t a l l e d in the m a c h i n e . V a r i a b l e c l a m p i n g force profiles can be p r o g r a m m e d a n d p r e d e t e r m i n e d with an e l e c t r o - h y d r a u l i c r e g u l a t o r . F o u r h y d r a u l i c p i s t o n s will be m o u n t e d on the t o p p l a t e o f the p r e s s in o r d e r to d i s t r i b u t e t h e c l a m p f o r c e m o r e u n i f o r m l y a n d to e n a b l e an o f f - c e n t e r force o p e r a t i o n . A l i n k a g e m e c h a n i s m p r e s s d r i v e n by an a.c. s e r v o - m o t o r with h i g h e r p o w e r will be a v a i l a b l e s h o r t l y in t h e m a r k e t . F a n u c C o m p a n y [14], p r o d u c e s a.c. s e r v o - m o t o r s o f up to 17 h p , with a r a t e d t o r q u e o f 147 N m. R e c e n t l y , it was r e p o r t e d that an a.c. s e r v o - m o t o r with u p to 70 h p c a p a c i t y was d e v e l o p e d for h i g h - p o w e r m o t i o n c o n t r o l a p p l i c a t i o n s [15]. W i t h i n c r e a s i n g s e r v o - m o t o r p o w e r , it will be p o s s i b l e to d e v e l o p special p r e s s e s with high t o r q u e a n d p r e c i s e s p e e d c o n t r o l o v e r a w i d e s p e e d r a n g e . T h e E R C / N S M is p l a n n i n g to p u r s u e t h e s e d e v e l o p m e n t s as p a r t o f a s t u d y to d e v e l o p n o v e l p r e s s c o n c e p t s a n d a p p l i c a t i o n s in m e t a l - f o r m i n g m a c h i n e tools. D e v e l o p m e n t o f the 30 t o n l i n k a g e d r i v e press, d r i v e n b y a.c. s e r v o - m o t o r s , is a significant s t e p in t h e Center's program. Acknowledgements--The authors appreciate the contributions of Jim Cardinal (CATIA simulations), Dinesh Damodaran (optimized rotary linkage drive force simulation), Darin Sherry (conceptual drawings of the press), Bruce Sellers (technical support for the 30 tons press) and Prof. Gary Kinzel for his support in modification of the KINLYN program for our needs. This project was supported by the National Science Foundation sponsored Engineering Research Center for Net Shape Manufacturing at the Ohio State University. The authors are especially grateful to Prof. Taylan Altan for his encouragement and suggestions during this research.
REFERENCES [1] S. YOSSIFONand T. ALTAN,Survey of Linkage Drive Mechanism Presses, Rep. No. ERC/NSM-B-9034. Engineering Research Center for Net Shape Manufacuring, The Ohio State University (1990). [2] J. A. PALE and R. Smvaum, Development of a computer controlled multi-action press for physical modeling of complex deformation, Proc. 18th North American Manufacturing Research Conf. Pennsylvania State University (1990). [3] S. YOSSIFON,D. MESSERLY,E. KROPP, R. SHIVPURIand T. ALTAN,A Servo-motor driven multi-action press for sheet metal forming, Int. J. Mach. Tools Manufact. 31, 345-359 (1991). [4] T. J. NYE and R. SHIVPURI,Investigation of an electric servo-motor driven linkage press for flexible forming: design synthesis, ASME Winter Annual Meeting, Dallas, U.S.A. (1990). [5] S. YOSSIFON,R. SHIVPURIand T. ALTAN, The a.c. servo-motor driven double toggle press: mechanism analysis and optimization, Rep. No. ERC/NSM-B-S-90-33. Engineering Research Center for Net Shape Manufacturing, The Ohio State University (1990). [6] Compumotor, Programmable Motion Control, Catalog No. 8000. Petaluma, California (1988). [7] S. YOSSIFONand R. SmvpuRI, Analysis and comparison of selected rotary linkage drives for mechanical presses, Int. J. Mach. Tools Manufact. 33, 175-192 (1993). [8] S. YOSS1FONand R. Smveum, Optimization of a double knuckle linkage drive with constant mechanical advantage for mechanical presses, Int. J. Mach. Tools Manufact. 33, 193-208 (1993). [9] C. H. REILLEY, C. A. MOUNT-CAMPBELL,D. J. GONSALVEZ,C. H. MARTIN, C. A. LEVIS and C. W. WAr,G, Broadcasting satellite service synthesis using gradient and cyclic coordinate search procedure, A1AA llth Communication Satellite Systems Conf., San Diego, CA (1986). [10] L. M. RETALLACK,Computer aided kinematic analysis of planar mechanisms containing higher and lower pairs, MS Thesis, The Ohio State University (1986). [11] CATIA (Computer Aided Three Dimensional Interactive Application), Kinematics Version 3 (5612011, 5613-011, 5612-111), Dassault System, Paris, France (1989). [12] PowerTrac Design Guide, Rolled Thread Ball Screws and Mounting Blocks, (brochure). Nook Industries, Cleveland, Ohio (1988). [13] A. G. ERDMANand G. N. SANDOR,Advanced Mechanism Design: Analysis and Synthesis, Vol. 1 and 2. Prentice Hall, Englewood Cliffs, NJ (1984). [14] Fanuc AC servo motor series, catalog No. ASV-12. Yamanashi Prefecture, Japan (1989). [15] W. ERICKSON, Intelligent drive for induction motors provides true servo accuracy in Information Quarterly, pp. 10-11. Mannesmann, Rexroth (1990).
DESIGN MOTOR
Vol. 33, No. 2, pp. 209-222, 1993.
0890-6955/9356.00 + .00 (~ 1993 Pergamon Press Ltd
CONSIDERATIONS FOR THE ELECTRIC SERVOD R I V E N 30 T O N D O U B L E K N U C K L E P R E S S F O R PRECISION FORMING S. YOSSIFONt and R. SHIveuPa$ (Received in final form 8 January 1992)
Abstract--This paper is the last part of a three part series on the design, analysis and construction of a flexible electric servo-motor controlled mechanical press for precision forming. In Part I, characteristics of the available mechanical drives for presses were compared and a double knuckle drive selected for this press application. This drive was shown to have the desirable characteristics of a large mechanical advantage and a constant ram velocity over the working stroke. Part II presented an enumeration search procedure for optimization of the link design for maximizing the mechanical advantage at the working stroke, with the constraint of a corLstant working velocity. This procedure was applied to the double knuckle linkage drive and the optimized dimensions selected. This paper presents the application of the above procedures to the development of a 30 ton double action press. Focus of this paper is on the scaling of the optimized design, force and torque analysis and the selection of press specifications. INTRODUCTION
MECHANICALpresses have been widely employed for the sheet and bulk metal-forming processes. However, due to their high velocities in the working zone, limited working strokes and poor controllability they have lost ground to hydraulic presses for cold forging and deep drawing applications. Recent attempts in the mechanical press design have focused on using a linkage drive mechanism to increase the approaching and return speeds, but to slow down the ram in the working zone. Another design objective is to maintain the specified load over a relatively long working stroke, while simultaneously reducing the crank torque and the frame strength requirements to achieve reductions in the weight, size and cost of the machine. Many different types of linkage drives proposed to achieve these objectives, have been studied [1]. Recently developed press drives include the double-toggle linkage, four-bar linkage, link type, drag crank linkage and the toggle lever presses. In all these drives, the idea was to satisfy the load-stroke characteristics for the specific metal-forming process. A 2.5 ton multiaction press driven by an a.c. servo-motor for precise position and velocity control was developed earlier at the Engineering Research Center (ERC) [2]. In this press, the angular displacement of the servo-motor was converted to a linear displacement by a worm-gear ball-screw jack. Recently, this press was modified for sheet metal-forming [3]. In parallel, two types of double-toggle linkage mechanisms were studied and proposed by Nye and Shivpuri [4]. In the first design, the motor and the screw were', held stationary as the linkage traveled through its stroke (type A in Fig. 1). The second arrangement eliminated the connector link by allowing the motor and the screw to pivot as the linkage traveled (type B in Fig. 1). A double-toggle rotary press with crankshaft (type C in Fig. 1) was proposed by Yossifon et al. [5]. The rotary drive (crankshaft) is capable of transmitting a higher torque (with the same motor) than the linear drive owing to its higher efficiency. Use of an a.c. servo-motor on the linkage press provides for the independent control of ram position and velocity. It can easily be interfaced with C N C machine controls because of its extensive use in the machine tool drives. The servo-motor is also capable of transmitting near constant torque over its speed range [6]. Control of the ram speed,
?Materials Division, NRCN, POB 9001 Beer-Sheva, Israel. SDepartment of Industrial and Systems Engineering, The Ohio State University, Columbus, OH 43210, U.S.A. 209
210
S. YOSSlFON and R. SHIVPURI
+Y +X
TypeA
Soot
+Y ~X
TypeB
÷Y f _~
-
A
D
' ~.__...
+X TypeC
B
Sout
FIG. 1. Schematic of type A, type B and type C design linkages.
independent of the other characteristics, provides more flexibility for the press user. The speed profile can be predetermined and programmed for each process to match the load-stroke characteristics of the press. Using a rotary transmission (crankshaft) driven by an a.c. servo-motor provides two load-stroke curves for each direction (clockwise or counterclockwise). In addition, the stroke length, bottom dead center (BDC) and top dead center (TDC) can be predetermined and programmed by rotating the crankshaft for only a portion of the cycle, and thereafter return back to the home position in the opposite direction. In the previous study [7], kinematic simulation programs were used for comparison between ERC rotary drive and four well-known rotary linkage drive mechanisms selected from the literature. The objective functions were the mechanical advantage and linkage drive size for the same total ram stroke. It was obtained from the kinematic
Design Considerations for the Double Knuckle Press
211
simulation that the mechanical advantage of the ERC rotary press is near constant along the major ram stroke. Therefore, the rotary double-toggle mechanism was chosen for prototype construction and optimization. For the verification of the kinematic simulation and design concepts, a wooden prototype of the press and a control unit were designed and built at the Center. The selected rotary double knuckle joint mechanism was then optimized based on design criteria and feasibility of the design [8]. Two objective functions (design criteria) were established: (a) the mechanical advantage of the mechanism; and (b) the size ratio (linkage dimensions non-dimensionalized with respect to the total stroke of the ram). In this study, a cyclic coordinate search method was implemented [9]. Several enumeration searches for the best arranged mechanism were completed. Based on these searches, an optimized rotary press design for a 5 in ram stroke was selected (Fig. 2). The goal of the current study (Part III) is to design an a.c. servo-motor driven 30 ton optimized double knuckle mechanical press. This press will be capable of controlled bulk forming experiments using model materials and sheet forming experiments such as thin sheet drawing. This press design procedure includes: rescaling and repositioning of the optimized rotary linkage drive for final comparison and selection; and force analysis of the optimized rotary linkage drive in order to evaluate the bearing forces and the required input links torque. From the analysis, preliminary construction design specifications are proposed. ANALYSIS OF THE LINKAGE DRIVE
Three different programs were used to study and compare the motion characteristics of the selected linkage drives. The first program, KINLYN [10], is capable of analyzing 2-D linkages for position, velocity and acceleration/deceleration. The second program, CATIA [11], is a general purpose CAD/CAM software for graphical display and analysis. The third program was a F O R T R A N code that read in the data file (ram stroke vs crank angle) and performed a finite difference incremental technique to find the derivative of the ram stroke vs input angle (in radians) curve for a rotary drive
Nod(
3
.5440,3.51~4)~Node 6
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2
~ x (- 1"7086'0"0000)
~
Link4
Node 1 (10.5607,-4.2968) Dimensions of the Optimized Press for Five Inch Stroke L2 L3 L4 L5 L6 L7
= 3.55917" = 8.82199" = 7.82290" = 7.52993" = 5.39344" = 8.55578"
Node 4
I
F~
FIG. 2. Schematic of the optimized ERC press for a 5 in ram stroke.
Node 7
212
S. YOSS1FON and R. SHIVPURI
and the derivative of the input stroke vs output stroke for a linear drive. The computer program calculated the necessary values, and generated a plot of mechanical advantage (MA) vs dimensionless stroke (h/H) curve for the linkage mechanism. Details on the use of these programs for kinematic analysis are presented in Parts I and II [7, 8] of this series. In our previous study [8], the optimal arrangement and the dimensions were obtained after several enumeration search steps for the ERC press (type C in Fig. 1) for a 5 in total stroke (see Fig. 2). The mechanical advantage improved from 2.5 to 4.0 due to the optimization of the linkage, this optimization also resulting in the size of the linkage mechanism being reduced for the same ram stroke. Details of this procedure were presented earlier [8]. Since the dimensions of the press are influenced by the arrangement of the links, it was decided to investigate the best location for the crankshaft center (CC) while keeping the link dimensions the same. In Fig. 2, the original location of the CC is at node 1, which is 17.2526 in to the left and 3.5254 in below the position of node 6 (the highest vertical point). As a first iteration, the CC was repositioned to a new location as in Fig. 3(a) (13.7 in to the left and 8.866 in below node 6) keeping the link dimensions and pivots the same. This new location substantially saves the space requirements. The output data file for the new design is plotted in Fig. 3(b). The total ram stroke obtained is about 3.5 in, instead of 5 in before repositioning. Therefore, this repositioning did not improve the motion performance of the optimized rotary linkage drive (type C). Further investigations of the location limits of CC were conducted using the linear arrangement (type A, Fig. 1), since it has similar characteristics to the rotary drive (type C, Fig. 1). Press A design was run on the CATIA program and the link positions along the entire ram stroke were plotted. Two positions of the angle of the linear input were selected as extremes of CC movement without changing the drive characteristics: 18.68 ° and 70° relative to the horizontal, Figs 4 and 5. Increasing the input angle significantly reduces the mechanism width which in turn reduces the overall size of the mechanism. However, the mechanical advantage of the "modified" linear press is near constant along the working stroke and too low, about 1.75 (Fig. 6).
FIG. 3(a). Links positions along the entire ram stroke of the repositioned optimized ERC rotary press.
Design Considerations for the Double Knuckle Press
213
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o
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o.
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C R A N K ANGLE (deg) Fro. 3(b). Ram stroke vs crank angle of the repositioned ERC rotary press.
18 ° 41'
i__
TDC
BDC
FIG. 4. The links positions of press A along the entire ram stroke. PRELIMINARY DESIGN OF THE LINKAGE DRIVE PRESS
The primary objective of the current study is to design and build a 30 ton press with the optimized double knuckle joint mechanism, driven by an a.c. servo-motor for accurate control of the velocity and the position of the ram. The following initial parameters were established for the press: (i) (ii) (iii) (iv)
servo-motor output torque of 240 in-lb; maximuJrn motor velocity is 1200 rpm; the total[ stroke of the ram is H = 5 in; and the required ram force is 20 tons at 1.5 in (h/H = 0.3) above the BDC.
With these parameters, the dimensional analysis of the optimized press had to be completed to compare it to the characteristics of the linear arrangement (type A). The specifications of the a.c. servo-motor chosen to power the 30 ton double-toggle
214
S. YOSSIFON and R. SHWPUR1
FXG. 5. The links positions of the repositioned press A along the entire ram stroke.
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8
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I 1
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FIG. 6. The mechanical advantage of press A and the repositioned press A vs ram stroke.
press are given in Table 1 [6]. This was the largest capacity a.c. servo-motor available in the U.S. at the time of design. The torque-speed characteristics of the servo-motor presents near constant torque capability over its speed range. As a result of this unique feature and the linearity of the linkage drive in the working zone, the ram force is proportional to the rotary mechanical advantage ( M A * ) of the press for a constant torque input. For the
Design Considerations for the Double Knuckle Press
215
TABLE 1. SPECIFICATIONS OF THE SERVO-MOTOR K H X - 7 4 0 FOR THE 30
TONPRESS[6] Continuous stall torque Peak torque Rated power Rated speed Rated motor stall current Rotor inertia Acceleration at peak torque Maximum resolution Motor weight
256.7 lb-in 513.0 lb-in 4.56 hp 1300 rpm 20 A 210.5 oz in-2 15,061 rad s-1 32,768 steps rev-1 58.3 lb
continuous duty region, the average output torque is 240 lb-in and the peak torque is limited to 513 lb'-in, A program was written that rescaled the linkage drives to match the required parameters and plotted desired curves of force and velocity. The stroke of the desired optimal press was chosen to be 5 in, a typical stroke needed for cold impact extrusion and sheet metal drawing of medium size components. After rescaling the lengths, the program performed a single position analysis to obtain the angle vs stroke data. The linear mechanical advantage data (MA) was calculated through the incremental slope method [7]. Knowing the necessary force and the desired mechanical advantage value at a certain point from the BDC, the necessary output torque to drive the linkage was calculated. The ram force F is determined by: F = (MA*) (M)/H
(1)
where M is the,, input link torque and H is the total stroke (5 in). The angular velocity of the input link was found by using the angular velocity of the servo-motor and dividing it by the gear ratio of the gearbox connected between the servo-motor and the linkage drive crankshaft. The gear ratio is calculated easily by determining the ram force required at the beginning of the working stroke. To find the linear velocity of the ram, the chain rule was used: dh dt
dh dot d a dt "
(2)
The ram velocity is dh/dt, and the value of det/dt is simply the angular velocity of the input link. The values of dh/dcL were calculated using an incremental slope method as presented in Ref. [7]. The rescaling of the linear drives is calculated through the following procedure. The ram force (Fout) is calculated as: Fo,t = (MA) (Fin)
.
(3)
The input force at the crank (Fin) is constant and is obtained through the following equation [12]:
Fin-
2'rrMi. L
(4)
where Min is the servo-motor output torque (constant torque) and L is the screw lead (in t u r n - l ) . To find the ram velocity, the chain rule from equation (2) had to be used. The ram velocity is dh/dt and the value of doddt is the angular velocity of the servo-motor. The value dh/dot is as follows:
216
S. YOSSIFON and R. SHIVPURI
dh da
L MA "
(5)
By substituting equation (5) into equation (2), the ram velocity obtained was:
L( o)
Vou,=M~ ~ •
(6)
The decision was made from design considerations, that at 1.5 in stroke from the BDC (nominal point) the required ram force is 20 tons. Knowing the mechanical advantage value over the total ram stroke, the output force and speed were predicted (using a.c. servo-motor type KHX-740 [6]). The values of force and velocity of the linear arrangement (type A, Fig. 1) and the repositioned linear press were plotted. One can see that for type A, the ram force curve at the working stroke rises up sharply toward BDC, and the ram velocity drops down non-constantly over its total stroke (Figs 7 and 8). For the repositioned type A, the ram force and the ram velocity curves were more acceptable. These results indicate that raising the CC increases the mechanical advantage but decreases the region of constant force and torque, while lowering CC does the opposite. The curves generated for the ERC optimized press are shown in Fig. 9(a) and (b) for ram force and ram speed respectively. The ram force at the working stroke is near constant and decreases moderately in the approaching stroke (Fig. 9(a)). Also the ram speed is near constant in the working stroke ( - 0 . 8 in s -1) and increases sharply at the approaching stroke for a constant input speed of 1200 rpm (Fig. 9(b)). These characteristics provide for (a) a near constant ram force and velocity in the working region and (b) rapid approach speed for reduced cycle time. Additional attempts to modify press C by repositioning the CC while using the same link lengths and geometry showed that press C with the location of Fig. 2 is the best design from the mechanical advantage, force and velocity characteristics, and the overall linkage size points of view.
8O
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Design Considerations for the Double Knuckle Press
217
4.2 Rep. Type A Type A 3.2
l
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After performing a kinematic analysis, optimization and rescaling of the doubletoggle rotary mechanism, a kinetostatic force analysis was done. This analysis took into account the inertial forces imposed on the links. The objective of the kinetostatic analysis is the de,termination of the bearing forces and the required input crank torque at the particular instant in the position being analyzed. Once analytical expressions are found for a single position, magnitudes of bearing forces and input torque for additional positions are easily determined by repeating the analysis for these additional positions. The matrix method of kinetostatic analysis was used to obtain the analytical expressions. This method considers all the inertial forces together, as compared to the superposition method wherein the effect of individual inertial forces are treated separately and then superimposed to determine their combined effect. One advantage of the matrix method is that the equations of motion are quickly derived. The most fundamental technique in force analysis is to break down a mechanism into several free bodies with compatible forces and torques applied to each free body. In each free body diagram of a link, the inertial forces are assumed to act at the center of mass and an inertial torque is added [13]. The mechanism is analyzed at each position by D'Alembert's principle: ~'~Fx + ( - M agx) = 0
(7a)
EFy + ( - ] ~ agy) = 0
(7b)
~,T + (-Igak) = O .
(7c)
In the above equations, ( - M agx) is known as inertial force component in the x direction, ( - M agy) is the inertial force in the y direction and ( - I g ak) is the inertial torque. The accelerations agx and agy a r e specified by subscript "g" denoting that these are accelerations of the centers of masses in the x or y directions. Based on the Cartesian coordinate system, the x, y components of forces, positions and accelerations are HTH 33:2-H
218
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Fro. 9. (a) R a m force curve vs ram stroke for the E R C optimized press. (b) Maximum ram velocity vs ram stroke for the E R C optimized press.
assumed positive in the positive x and y directions, respectively• The counterclockwise direction is assumed as the positive torque direction• For each link, one can write three equilibrium equations (equations (7a), (7b) and (7c)), and for each free node (node numbers 2, 3 and 5 in Fig. 2) two equilibrium force equations (for the x and y directions)• Thus, there are 24 equations and 24 unknowns, and this system of equations can be solved• Since the positions of the nodes are known from the kinematic analysis, the system of equations is linear in the unknowns and can be expressed in the matrix form:
[L][FB] = [FI]
(8)
where [L] is the square matrix of known linkage parameters, [Fa] is the column matrix of unknown joint forces and input torque, and [F~] is the column matrix of known external loads plus inertial forces and torques. The solutions were solved using a linear equation solver by implementation of the Gauss elimination method• From design considerations, the linkage strength was to be calculated for a maximum ram force of 50 tons (approximately the maximum possible
Design Considerations for the Double Knuckle Press
219
force generated by the linkage in its entire stroke). These equations were then solved assuming a con,;tant ram force of 50 tons throughout the working stroke, and the torque, which was to be supplied by the gearbox unit through the servo-motor, was checked. However, since the servo-motor integrated with a transmission unit of 205 gear ratio did not have the capacity to supply torque over its rated maximum torque capacity, which is limited to 41.0 ton-in (400 lb-in motor torque) by a special adjustable clutch (the continuous permissible crank torque is limited to 24.6 ton-in), it would be able to supply only that maximum torque and the ram force that would be available would drop below its 50 tons capacity. The factor by which the servo-motor's rated torque capacity fell short of the torque requirement imposed by the constant ram force of 50 tons was used to accordingly reduce all the forces. Thus, the working stroke was simulated as two load steps. Near the BDC, links 6 and 7 (Fig. 2) were nearly vertical and this allowed a large fraction of the full load to be transmitted through the links to the ground. However, as the links moved away from the BDC, the mechanical advantage of the mechanism dropped and this led to very high torque requirements on the servomotor. Now the., condition used to solve the equations was that of constant torque supply by the servo-motor. Hence, the working region was divided into two regions, one where the ram force would be a 50 tons constant (curve "a" in Fig. 10) and the other when the torque transmitted by the servo-motor and the gear unit was a 41 tonin constant (curve "b" in Fig. 10). Figure 11 shows the acting load in each individual link vs the crank angle along the working stroke. One can see that links 3, 5, 6 and 7 are subjected to compression forces along all of the working stroke, compared to link 4 which is subjected to tensile forces all over the working stroke. However, link 2 is subjected to tensile loads during a part of the stroke and compression loads during the rest of the stroke. The maximum link loads are presented in Table 2. These link loads will help determine the dimensions of the links and the joints of the drive linkage. In the detailed design of the joints, rigidity and wear considerations often override the strength requirements.
j-
50
B
/
4O
/
/
/
50
I
40
'7 30
30
ia_ j-.
~
20
20
10
o 260
\
I
I
I
I
280
300
320
340
10
0 360
CRANK ANGLE, 0 (deg) FIG. 10. Plots of (a) input torque vs crank angle and (b) ram force vs crank angle, produced along the working stroke, when the maximum input torque is 41 ton-in and maximum ram force is 50 tons.
220
S. YOSSIFON and R. SHIVPURI 25 LINK 4
20
i0 ~ . ~ L I N K
-10
2
~
3
_~ -2o -
~
-30
-5(
260
280
300
320
340
360
CRANK ANGLE, e (deg)
FIG. l l . Link load vs crank angle in the working region for the linkage drive scheme shown in Fig. 2. TABLE 2. M A X I M U M LINK LOADS ALONG THE WORKING STROKE FOR MAXIMUM 5 0 TONS RAM STROKE
Link No.
2
Max. load (tons)
3
- 9 . 0 -14.7
4
5
6
7
24
-30
-50
-50
DESIGN SPECIFICATIONS FOR THE 30 TON PRESS
Based on the comparison and analysis [7] and the linkage optimization [8], a double action 30 ton capacity press is being built at the Engineering Research Center. The press uses a double knuckle joint linkage arrangement to drive the top punch with an a.c. servo-motor. The second action (clamping plate), is driven by four hydraulic pistons mounted on the top plate. The press will be used for precision bulk and sheet metalforming. The press consists of four circular guide posts, connected between the top plate and the base plate. The clamping plate travels with controlled velocity along the guide posts on bushings. The ram travels along the press center and is guided by bushing connected to the top plate. The angular displacement of the a.c. servo-motor is transferred to the ram through a reduction transmission and a double knuckle joint mechanism. The transmission system is mounted on the top plate. The conceptual drawing of the press construction is shown in Fig. 12. The focus of this paper being the mechanical drive, the details of the hydraulic drive for the clamping plate and the computer control system are not included. The planned press specifications and the performance are given in Table 3. CONCLUSIONS AND FUTURE WORK
In the current study it was proved theoretically that the orginal optimized rotary drive mechanism can be adopted for the 30 ton press design and construction. From kinematic and kinetostatic analysis, the motion characteristics and link loads were predicted for a KHX-740 servo-motor drive [6], which is available in the market. The designed press has a maximum actual ram force of 30 tons and maximum ram speed of 0.8 in s -1 at the working stroke.
Design Considerations for the Double Knuckle Press
221
Input from servomotor
or1 for linkage
G
loggle linkage
Rc,tar!
r bearing
)er platen Clan h guidance
I=ou
or lower punch
r platen
frame
FIG. 12. A schematic drawing of the 30 ton double knuckle joint press showing the mechanical drive components (side view). TABLE3.
PLANNED PRESS SPECIFICATIONS AND PERFORMANCE
Maximum ram force: Maximum clamping force: Maximum ram stroke: Ram fi~rce at 1.5 in stroke from BDC: Maximum clamping stroke: Maximum pistons working pressure:
30 tons 25 tons 5 in 20 tons 12 in 2000 psi
Maximum down stroke clamping velocity: Maximum up stroke clamping velocity: Maximum working ram velocity: Maximum ram velocity: Ram velocity profile:
0.55 in s 0.45 in s -~ 0.8 in s -~ 2.5 in s User specified/PC control
Maximum press height: Maximum press width: Maximum press depth: Minimam shut height: Daylight: Effective die set space:
~ 110 in ~ 50 in ~ 42 in 12 in 24 in 28 × 21 in
D e t a i l e d design o f th e 30 t o n press will be m a d e in t h e n e a r f u t u r e . T h e specifications o f t h e press witlh c o n c e p t u a l d r a w i n g s w e r e p r e s e n t e d earlier. T h e m o d i f i c a t i o n s o f a s e r v o - m o t o r d r i v e n m u l t i a c t i o n press f o r s h e e t m e t a l - f o r m i n g will b e i m p l e m e n t e d in t h e 30 t o n press design. M e a s u r e m e n t d e v i c e s such as a p u n c h l o a d d e v i c e , p r e s s u r e t r a n s d u c e r f o r c l a m p i n g f o r c e m e a s u r e m e n t an d a d e v i c e f o r m e a s u r i n g t h e p u n c h
222
S. YOSSIFONand R. SHIVPURI
d i s p l a c e m e n t will b e i n s t a l l e d in the m a c h i n e . V a r i a b l e c l a m p i n g force profiles can be p r o g r a m m e d a n d p r e d e t e r m i n e d with an e l e c t r o - h y d r a u l i c r e g u l a t o r . F o u r h y d r a u l i c p i s t o n s will be m o u n t e d on the t o p p l a t e o f the p r e s s in o r d e r to d i s t r i b u t e t h e c l a m p f o r c e m o r e u n i f o r m l y a n d to e n a b l e an o f f - c e n t e r force o p e r a t i o n . A l i n k a g e m e c h a n i s m p r e s s d r i v e n by an a.c. s e r v o - m o t o r with h i g h e r p o w e r will be a v a i l a b l e s h o r t l y in t h e m a r k e t . F a n u c C o m p a n y [14], p r o d u c e s a.c. s e r v o - m o t o r s o f up to 17 h p , with a r a t e d t o r q u e o f 147 N m. R e c e n t l y , it was r e p o r t e d that an a.c. s e r v o - m o t o r with u p to 70 h p c a p a c i t y was d e v e l o p e d for h i g h - p o w e r m o t i o n c o n t r o l a p p l i c a t i o n s [15]. W i t h i n c r e a s i n g s e r v o - m o t o r p o w e r , it will be p o s s i b l e to d e v e l o p special p r e s s e s with high t o r q u e a n d p r e c i s e s p e e d c o n t r o l o v e r a w i d e s p e e d r a n g e . T h e E R C / N S M is p l a n n i n g to p u r s u e t h e s e d e v e l o p m e n t s as p a r t o f a s t u d y to d e v e l o p n o v e l p r e s s c o n c e p t s a n d a p p l i c a t i o n s in m e t a l - f o r m i n g m a c h i n e tools. D e v e l o p m e n t o f the 30 t o n l i n k a g e d r i v e press, d r i v e n b y a.c. s e r v o - m o t o r s , is a significant s t e p in t h e Center's program. Acknowledgements--The authors appreciate the contributions of Jim Cardinal (CATIA simulations), Dinesh Damodaran (optimized rotary linkage drive force simulation), Darin Sherry (conceptual drawings of the press), Bruce Sellers (technical support for the 30 tons press) and Prof. Gary Kinzel for his support in modification of the KINLYN program for our needs. This project was supported by the National Science Foundation sponsored Engineering Research Center for Net Shape Manufacturing at the Ohio State University. The authors are especially grateful to Prof. Taylan Altan for his encouragement and suggestions during this research.
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