926 Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942
Journal of Zhejiang University-SCIENCE A (Applied Physics & Engineering)
ISSN 1673-565X (Print); ISSN 1862-1775 (Online)
https://www.doczj.com/doc/fc6319803.html,/jzus; https://www.doczj.com/doc/fc6319803.html, E-
mail: jzus@https://www.doczj.com/doc/fc6319803.html,
Quantitative measures for assessment of the hydraulic
excavator digging efficiency*
Dragoslav JANOSEVIC?1, Rosen MITREV2, Boban ANDJELKOVIC1, Plamen PETROV2
(1Faculty of Mechanical Engineering, University of Nis, Nis 18000, Serbia)
(2Faculty of Mechanical Engineering, Technical University of Sofia, Sofia, Bulgaria)
?E-mail: janos@masfak.ni.ac.rs
Received Nov. 18, 2011; Revision accepted Mar. 27, 2012; Crosschecked Nov. 1, 2012
Abstract: In this paper, quantitative measures for the assessment of the hydraulic excavator digging efficiency are proposed and developed. The following factors are considered: (a) boundary digging forces allowed for by the stability of an excavator, (b) boundary digging forces enabled by the driving mechanisms of the excavator, (c) factors taking into consideration the digging position in the working range of an excavator, and (d) sign and direction of potential digging resistive force. A corrected digging force is defined and a mathematical model of kinematic chain and drive mechanisms of a five-member excavator configuration was developed comprising: an undercarriage, a rotational platform and an attachment with boom, stick, and bucket. On the basis of the mathematical model of the excavator, software was developed for computation and detailed analysis of the digging forces in the entire workspace of the excavator. By using the developed software, the analysis of boundary digging forces is conducted and the corrected digging force is determined for two models of hydraulic excavators of the same mass (around 17 000 kg) with identical kinematic chain parameters but with different parameters of manipulator driving mechanisms. The results of the analy- sis show that the proposed set of quantitative measures can be used for assessment of the digging efficiency of existing excava- tor models and to serve as an optimization criterion in the synthesis of manipulator driving mechanisms of new excavator models.
Key words: Hydraulic excavators, Digging efficiency, Quantitative measures
doi:10.1631/jzus.A1100318 Document code: A CLC number: TV53
1 Introduction
The hydraulic excavators are popular multifunc-tional construction and mining machines. The main function of the hydraulic excavators of all types and sizes is the cyclic digging and transfer of soil. This function is achieved by use of an open kinematic chain consisting of undercarriage L1, an upper struc-ture L2 and a front attachment with boom L3, stick L4 and work tool L5 (Fig. 1). For digging operations below the ground level, the toward oneself technol- ogy (in relation to the excavator operator) is em- ployed and a backhoe attachment is used (Fig. 1a). * Project (No. 035049) partly supported by the Ministry of Education and Science of the Republic of Serbia
? Zhejiang University and Springer-Verlag Berlin Heidelberg 2012 For digging operations above the ground level the away from oneself technology and a shovel attach- ment are used (Fig. 1b).
During digging operations, the occurring dig-ging resistive force acting against the manipulator tool is overcome by digging forces. The digging forces of an excavator are exerted by the kinematic chain of the excavator, which is powered by the fol-lowing driving mechanisms: (1) hydraulic motors for motion of swing and travelling bodies; (2) double acting hydraulic cylinders for powering of the ma-nipulator links.
For an optimal design of the front manipulator and driving mechanisms, it is necessary to perform a detailed analysis of the digging forces and the digging resistive forces in the entire workspace of the excava- tor. The conducted analytical and experimental
Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942 927
L
research that points out the importance of knowledge about the digging resistive forces for development and analysis of hydraulic excavators is related to: (1) analytical modeling and experimental determination of the value and the change in the digging resistive force during the excavation process (Maciejewski and Jarzebowski, 2002; Maciejewski et al ., 2003; Yang et al ., 2008); (2) development of mathematical models for kinematic and dynamic analysis of exca- vators (Budny et al., 2003; Towarek, 2003; Hall and McAree, 2005; Gu et al ., 2007); (3) development of control systems for automation of the digging proc- ess (Plonecki et al ., 1998; Ha et al ., 2000; Chang and Lee, 2002; Lee and Chang, 2002; Flores et al ., 2007; Lin et al ., 2008); and, (4) definition of quantitative measures for analysis and assessment of excavator digging efficiency in the workspace.
L 3
2 L 4
L 5
(a)
L 3
L 2
L 4
L 5
(b)
Fig. 1 Hydraulic excavators with backhoe (a) and shovel attachment (b)
As indicated by the manufacturer, stick and bucket digging forces are important parameters of the excavator. They are defined by appropriate stan- dards (ISO 6015, 2006; SAE J1179, 2008) as one of the characteristics of the excavator digging function. For robotic manipulators, quantitative measures of workspace attributes are defined, which includes structural length index and manipulability measure
(Craig, 2005). Some of these measures can be used for the assessment of the dynamic performance of shovels (Lipsett, 2009). In many cases, the above mentioned measures are not sufficient for assessment of digging possibilities and efficiency of the excava- tor in the workspace.
In this paper, a set of quantitative measures for assessment of the hydraulic excavator digging effi- ciency is proposed and developed.
For a comprehensive analysis of values and di- rections of digging forces in a specific position of the front manipulator, a hodograph of boundary digging forces at the top of the bucket tooth is defined. It has a form of a polygon whose sides are composed by vectors of boundary forces exerted by the manipula- tor driving mechanisms and vectors of boundary forces allowed by the stability of the excavator. The ratio of the computed digging force and the potential digging force is defined as a measure of the digging efficiency. The computed digging force is deter- mined by a mathematical model of the excavator. Computations are performed for a given position and orientation of the manipulator and pressures in the hydraulic cylinders of driving mechanisms. The po- tential digging force represents the minimum value of boundary digging forces (Dudczak, 1977).
A hodograph of the effective digging forces is defined as a part of the hodograph of the boundary digging forces, for which the dot product of the dig- ging velocity and digging force vector is positive. The ratio of the area and range of the effective dig- ging forces hodograph is accepted as a criterion for excavator digging efficiency, whereas the area of the effective hodograph represents the mean value of the allowed digging forces. The range of the hodograph of the effective digging forces reflects the degree of compatibility of manipulator actuator parameters, excavator weight distribution, and digging force dis- tribution (Budny, 1989).
A corrected digging force is set as a measure of digging efficiency in the entire workspace. The fol- lowing considerations are taken into account: bound- ary digging forces allowed by the stability of an ex- cavator; boundary forces exerted by the driving mechanisms of an excavator; and factors which re- late to the digging position of the manipulator in the workspace as well as the sign and direction of the potential digging resistance action (Janosevic, 1997).
928 Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942
In this paper, mathematical models of the exca-vator kinematic chain, driving mechanisms and boundary digging forces are proposed and developed.
2 Mathematical modeling of excavator
A side view of the excavator is shown in Fig. 2. The mathematical model of the excavator consists of the mathematical model of the excavator kine-matic chain and that of the manipulator driving mechanisms.
Model of the excavator kinematic chain
The excavator represents a five-link open kine-matic chain which consists of the travelling body L1, the swing body (rotational platform) L2 and the three-link front manipulator consisting of a boom L3, a stick L4, and a bucket L5 (Fig. 2).
The links of the front attachment kinematic chain are connected by kinematic pairs of the fifth class–rotational joints with a 1 DOF with different orientations in the space. The kinematic chain of the front manipulator which is a part of the excavator model is planar. The centers of the manipulator joints O i (i=3, 4, 5) are penetration points of joint axis through the plane of symmetry of manipulator chain tor is considered as an open kinematic chain with a last link (bucket) that is subjected to the digging re-sistive force W (Janosevic, 1997).
To describe mathematically the kinematic model of the excavator, the following coordinate systems have been assigned:
(1)Global reference frame OXYZ with unit vectors i, j, k along the coordinate axes OX, OY, and OZ, respectively. The support base lies in the hori-zontal plane OXZ and the vertical axis OY coincides with the axis of the rotational joint between the un- dercarriage and upperstructure O2.
(2)Body fixed coordinate systems O i x i y i z i (i=1, 2, 3, 4, 5), which are connected to each link L i of the kinematic chain. The coordinate systems beginning O i is situated in the center of joint by which the chain member L i is connected to the previous member L i?1. The bucket is connected to a coordinate system whose axis O5x5 passes through the center of the joint O5 and the top of the cutting edge of the bucket O w.
In the case of the stationary undercarriage, the coordinate system O1x1y1z1 coincides with the global coordinate system.
The geometry of the kinematic chain links L i is defined in its local coordinate system O i x i y i z i by a set of geometrical parameters:
links.
Joint axes O i (i=3, 4, 5) are parallel to each other, and the centers of joints lie in the same plane–
L
i
= {e i , s i , t i , m i },
(1)
the plane of the manipulator. The penetration of the where e i = {e ix , e iy , e iz } is a unit vector (ort) of the
cutting edge of the bucket through the plane of the manipulator represents the center of the cutting edge of the bucket O w.
The adopted assumptions of the mathematical model of the excavator kinematic chain are: (1) The links of the excavator kinematic chain are modeled as rigid bodies; (2) The first joint between the travel body and the ground is a rotational joint O11 (or O12), whose axes represent the lines along which the exca- vator could tip over; (3) During the working cycle, the excavator is stable and there is no motion in the first joint; (4) During the working cycle, the follow- ing forces act on the excavator: digging resistive force W, weight of kinematic chain and driving sys- tem links, weight of soil in the bucket; (5) During the digging operation, the kinematic chain of the excava- axes of joint O i used to connect links L i and L i?1; s i={s ix, s iy, s is} is the vector of positions of the center of joint O i+1 used to connect links L i and L i+1, whereas the magnitude of vector s i represents the kinematic length of member L i; and t i={t ix, t iy, t iz} is the vector of position of the center of mass m i of link L i.
To define the digging forces, the planar position of the kinematic chain configuration of an excavator is observed when potential (widthwise x-x or lengthwise z-z) lines along which an excavator can tip over are parallel to the axes of manipulator joints, where it is assumed that the centres of joints and masses of all members of the kinematic chain of an excavator lie in the vertical plane OXY of the abso- lute coordinate system of the model.
Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942
929
3
M p3
O 3
B 2
L 3
t 3
C 3
m c3 r c4
x 2
m 3 A 3
r 4
m c4
C 4 A 4
M p4
x 2B 5
O 4
θ4 m 4 r t 4 m 1
L 4
O 45
m c6
m c5 C 5
r 11
O 1
1
x 1 HW 4 r w
HW 12
W 4min
L 5
W 11min
W 11min M p5 O 5
O w
A 55 m 5
W 5
A m c7
z x
HW 3
HW 5
W 3min
W 5min
W 4
θ5
x 5
x 4
O 3 x 1 m
O 4
m 4 O 5
5 e
3
m
x 2
x 5
e
e 4 e 5
3
e 1
1
z 3
z
x
(b)
HF e
z 4 z 5
φ5
n
O 45 A 5 = a 5 , O 45B 5 = b 5 , O 5O 45 = b 45 , O 5 A 55 = a 55 , A 5 A 55 = c 55 .
F O w
v
φ
Fig. 2 Side view of a hydraulic excavator with a backhoe attachment
(a) Boundary digging resistive forces; (b) Hodographs of potential and effective digging resistive forces
Mathematical model of the manipulator driv- ing mechanisms
Тhe functions of the excavator are performed by the kinematic chain links powered by driving mechanisms. The front manipulator driving mecha- nisms are powered by double action hydraulic cylin- ders connected directly or through a lever system to the links L i ?1 and L i .
Hydraulic cylinders of the driving mechanisms of the boom C 3 and the stick C 4 are directly con- nected to the links. Cylinder of the driving mecha- nism of the bucket C 5 is connected to the bucket through a four-bar mechanism.
In the mathematical model of manipulator driv- ing mechanisms, the following assumptions are made: (1) the position of the center of mass of cylinders is
Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942931
located in the middle of the current length of the cyl- nism (Fig. 2); b
5 = {b
5 x
, b
5 y
} is the coordinates of
inder; (2) the masses of the joint elements and trans- mission levers of driving mechanisms are added to the links of manipulator kinematic chain; (3) friction in the driving mechanisms joints, friction in the cylinders and the influence of pressure in return pipes of cylin-ders are neglected.
The driving mechanism C i of the manipulator is determined by the set of parameters: the position of the center of joints in which a bucket hydro cylinder is connected to the manipulator stick; c55 is the length of the link in the transmission part of the driving mechanism; a55={a55x, a55y} and b45= {b45x, b45y} are coordinates of the position of the cen-ter of joints in which transmission levers are con- nected to the links; m c6 is the mass of the lever of the bucket cylinder in the transmission part of the driv- ing mechanism; and m c7 is the mass of the link in the
C
i = C
c i
C
p i
, ? i = 3,4,5, (2) transmission part of the driving mechanism.
Geometrical quantities of the model
where C c i is denoted as the set of parameters of the
hydraulic cylinders, and C p i is the set of transmission parameters of the driving mechanism.
The parameters of the cylinders are determined by the following set:
The generalized coordinates of the excavator kinematic chain are represented by the relative an- gles θi (Fig. 2) between axes of two consecutively connected links L i and L i?1. The change of the length c i of the hydraulic cylinders in the interval of its boundary values c i=[c i p, c i k] leads to the change of
C
c i = {
d i1 , d i 2 , c i p , c i k , m c i , n c i },? i = 3,4,5, (3) th
e generalized coordinates θ
i
in the interval θi=[θi p,
θi k], where θi p is the starting and θi k is the final angle
where d i1 and d i2 are the piston and rod diameters, respectively; c i p is the minimum length of the cylin-der with the connecting rod completely pulled in; c i k is the maximum length of a hydro cylinder with the connecting rod completely pulled out; m c i is the mass of cylinder; and n c i is the number of cylinders in the of the relative position of link L i in relation to the previous link L i?1.
The relative angle θi o between the links L i and L i?1 is expressed as follows:
driving mechanism.
The set of transmission parameters of the driv- ing mechanisms of the boom C3 and stick C4 are de-
θ
i o
= θ
i k
- θ
i p
.
The position of link L i
(6)
in relation to the hori-
termined by the set:
C
p i = {a
i
, b
i
}, ?i = 3, 4, (4)
zontal OXZ plane of the absolute coordinate system
is determined by the angle:
i
where a
i = {a
ix
, a
iy
}, b
i
= {b
ix
, b
iy
} are the coordi-
φ
i
= ∑θi , ?i = 3, 4, 5.
i =3
(7)
nates in the local coordinate systems of the position of the center of joints in which hydraulic cylinders are connected to the links of the driving mechanism (Fig. 2).
The subset of transmission parameters of the driving mechanism of the bucket C5 is determined by
Unit vector e i of the joint O i axis, vector r i of the center of joint O i, vector r w of the centre of the bucket cutting edge and vector r ii of the center of mass of link L i in the absolute coordinate system are determined by Eqs. (8)–(11), respectively:
the set: e A
, (8)
C
p5 = { a
5
, b
5
, c
55
, a
55
, b
45
, m
c6
, m
c7
}, (5)
i
=
i o
e
i
,
i -1
r
i
= ∑ A i o s i ,
i =1
5
?i = 2,3, 4, 5, (9)
where a5 is the length of the lever of the bucket cyl- inder in the transmission part of the driving mecha- r
w
= ∑ A iοs i ,
i =1
(10)
930 Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942
1 0 0 A 1o = 0
1
0 .
(12)
0 0 1
1 0 0 A 2o = 0
1 0 . 0
1
?
r t i = r i + A i o t i ,
(11)
for certain axes of joints O i (i =3,4,5).
Moments M oc i of gravitational forces of the where A i o is a transformation matrix, used to transfer the vectors from the local coordinate system O i x i y i z i to the absolute coordinate system OXYZ . The ele- ments of the transformation matrices A i o for the con- links of excavator driving mechanisms for certain axes of joints O i (i =3,4,5) are determined by the fol- lowing equation (Fig. 2):
sidered planar position of the excavator kinematic ?
M = -g n c3 m c 3 ((r - r ) ? j ) ? e ,
chain are given as follows.
1. Transformation matrix A 1o from the local ? oc3
? ?
2 k =7 A
3 3 3 system O 1x 1y 1z 1 between absolute system OXYZ (Fig. 2):
?
- g ∑ n c k m c k ((r c tk - r 3 ) ? j ) ? e 3
?
k =4
?
?i = 3, ? ?M = -g n c4 m c4 ((r - r ) ? j ) ? e M oc i = ? ? oc4 2
k =7 A 4 4 4 (16) ? - g ∑ n c k m c k ((r ctk - r 4 ) ? j ) ? e 4 , ? k =5 2. Transformation matrix A 2o from the local sys- ? ?i = 4, ? ?
M = -g m c7 r - r ? j ? e
tem O 2 x 2 y 2 z 2 between the absolute system OXYZ :
? oc5
?
(( A 55 5 ) ) 5 , 2 (13)
?
?
where r c tk
?i = 5, is the vector of the position of center of
3. Transformation matrix A i o from the local sys- tem O i x i y i z i (i =3,4,5) between the absolute system OXYZ :
mass of hydro cylinders and the levers of the trans- mission part of driving mechanisms, where it is as- sumed that the centers of mass of hydro cylinders and the levers of the transmission part of driving mechanisms are at the middle of their kinematic cos φi -sin φi 0
lengths c i , a 5 and c 55 (Fig. 2a); r A 3 , r A 4 are vectors of A i o = sin φi cos φi 0 , ? i = 3, 4,5. (14)
the position of centers of joints A 3 and A 4 where hy- 0 0 1
dro cylinders are linked to the stick and boom of the attachment; r A 55 is the vector of the position of cen- Mechanical parameters of the model
The moment of gravitational forces of the links
in the kinematic chain as well as the links of the ex- cavator driving mechanisms for certain axes of joints O i (i =3,4,5), when the bucket is empty, is determined by (Fig. 2)
ter of joint A 55 where the link of the transmission part of the mechanism is connected to the attachment bucket (Fig. 2a).
Gravitational moments for the first joint of the kinematic chain (i =1, with respect to x-x or z-z tip over line) are:
k =5
M o i = -g ∑ m k ((r tk - r i ) ? j ) ? e i + M oc i ,
(15)
?
? o11
k =5 ∑ k tk 11 1 k =i
M = -g m ((r k =1
k =7
- r ) ? j ) ? e
? i = 3, 4, 5,
? - g ∑m c k ((r c tk - r 11 ) ? j ) ? e 1 , M = ? k =3 (17) o1 ? k =5
?M o 12 = -g ∑ m k (r tk - r 12 ) ? j ) ? e 1 where e i = {0,0,1} is the unit vector of axes of joints ? k =1 k =7 O i (i =3,4,5), and M oc i is the moments of gravitational forces of the links of excavator driving mechanisms
?
- g m
?
∑ c k ? k =3
((r c tk - r 12 ) ? j ) ? e 1 ,
932 Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942
i = 3 4 5 3 4 5 ?
where e 1 = {0, 0,1} is the unit vector of the first rota- tional joint (for x-x or z-z tip over line), and r 11 , r 12 are vectors of the center of the appropriate first rota- tional joint.
The moments of gravitational forces of the links of the kinematic chain, driving mechanisms links and the weight of the soil in the bucket are related by
3 Digging forces
The vector of digging force F lies in the plane of the manipulator and acts in the center of the bucket cutting edge. It has the same magnitude, di- rection and opposite sign to the vector of digging resistive force W :
M zi = M o i - gm z ((r t 5 - r i ) ? j ) ? e i ,
? i = 1, 3,4,5,
F = - W .
(21)
(18)
where m z is the mass of the soil in the bucket. It is assumed that the center of the mass of the soil coin- cides with that of the bucket.
The direction of digging resistive force W with re- spect to the horizontal OXZ plane is determined by the angle:
5
Depending on the position of the bucket, the mass of the soil in the bucket is determined by the expression:
φw = ∑θi + θw ,
i =3
(22)
?ρz ?V |cos φ5 |, m z = ? ?0,
? 270? ≥ φ5 ≥ 90?, ? 270? < φ5
< 90?,
(19)
where θw is the angle between the direction of the digging resistive force and O 5x 5 axis of the local co- ordinate system of the bucket L 5.
The direction and sign of digging resistive force W in the absolute coordinate system are determined
where ρz is the density of the soil, V is the capacity of the bucket, and φ5 is the angle of the position of the bucket in the absolute coordinate system (Fig. 2b).
Pressure p and flow Q of the hydrostatic exca- vator system are transformed according to the trans- mission function of driving mechanisms into angular
by the unit vector ort W :
ort W = i cos φw + j sin φw .
Boundary digging resistive forces
(23)
speed θ and driving moment M p i of the link. The The value of the digging resistive force is lim- ited by the conditions of excavator tip over and maximum driving moments M p i max of the manipula- tor mechanisms in both directions of action are: maximum forces exerted by the driving mechanisms. The boundary digging resistive force limited by the excavator stability is determined by the equations of ?
M = sign (θ ) ? r ? n d 2 π i 1
p , static equilibrium with respect to the first joint of the ? p i 1max ?
i c i c i 4 max excavator kinematic chain. They are represented by ? M p i max ? ?i = 3, 4,5, θ > 0, θ < 0, θ < 0, (d 2 - d 2
) π the values of two boundary digging resistive forces.
The first boundary digging resistive force W 1x
?M = sign(θ ) ? r ? n i 1 i 2 p ,
(Fig. 2) is limited by the friction force between the ? p i 2max i c i c i 4
max ?? ?i = 3, 4,5, θ < 0, θ > 0, θ > 0,
(20) undercarriage and the ground as follows:
mg ? μ where r c i is the transmission function of the driving mechanisms which depends on the length of a hydrau- W 1x =
p
, cos φw
(24)
lic cylinder, the length of transmission levers and co- ordinates of the position of the joints of the hydraulic cylinders (Janosevic, 1997); and p max is the maximum pressure of the hydrostatic excavator system.
where m is the total mass of the excavator; μp is the dry friction coefficient between the undercarriage and the ground. The minimum value of this bound- ary digging resistive force W 1x min is
Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942 933
= = W
W 1x min = mg ? μp .
(25)
tion of the excavator kinematic chain along with the
maximum driving moments M p i max , are determined The second boundary digging resistive force W 1o is determined from the excavator tip over condi- tions. Depending on the position of the excavator kinematic chain and ort W , the boundary digging re- from the equations of the static equilibrium with re- spect to the axes of manipulator joints O i (i =3,4,5). These are given by (Fig. 2)
sistance W 1o limited by the static excavator stability is determined from the equations of equilibrium with respect to one of the rotational joints O 11 and O 12, W i =
-M p i max - M o i
((r w - r i ) ? ort W ) ? e i , ? i =3,4,5.
(29)
whose axes represent the potential tip over lines of the excavator as follows (Fig. 2):
The minimum values of boundary digging resis-
tive forces W i min which can be overcome by the driv- ing mechanisms are obtained:
? ?W 11 ((r -M o11 ,
- r ) ? ort W ) ? e
-M
- M
?
w 11 1 p i max
o i
? ? y > 0, φ > φ > φ + 180o ,
W i min =
,
| r - r | (30)
? w 12 w 11 w i
? ? y < 0, φ > φ > φ - 180o , ? (r - r ) ⊥ ort W , i = 3,4,5.
W = ?
w 11 w 12
(26)
w
i
1o ? -M ?W = o12 ,
? 12 ((r - r ) ? ort W ) ? e Hodographs of digging resistive forces ? w 12 1
? ? y > 0, φ + 180o > φ > φ ,
For a specified position of the kinematic chain
? w 12 w 11
links, the hodograph of the boundary resistive force ? ? y < 0, φ - 180o > φ > φ , ? w 11 w 12
where y w is coordinate of the bucket top, and φ11 , φ12 are determined by
φ = arccos ? (r w - r 11 ) ? i ? ,
HW i (it is equivalent to the hodograph of the digging force HF i ) (Fig. 2a) is a line, perpendicular to the vector of the boundary resistive force W i min passing through the point O w . The value of the force W i min is the minimum value, determined from the operation of the driving mechanism in both directions 11 r - r (Dudczak, 1977; Janosevic, 1997).
φ ? w 11 ? ? (r - r ) ? i ? = arccos w 12 ,
(27)
Boundary resistive force W g for the known ort W and the position of the kinematic chain consists of a 12 r - r
? set determined by certain boundary digging resistive ? w 12 ?
forces (or digging forces):
where (r w - r 11 ) OX axes.
and (r w - r 12 ) are vectors built with | -F g |= W g = {W 1o ,W 1x ,W 3 ,W 4 ,W 5 }.
(31) The minimum values of the boundary digging resistive forces W 1min limited by the static excavator stability are obtained from
Potential resistive force W m (or digging force F m ) is the value of the minimum resistive force from the set
of boundary digging resistive forces:
?
-M o11 , ?(r - r ) ⊥ ort W , ? 11min r - r w 11 = ? w 11 -F m = W m = min{W 1o ,W 1x ,W 3 ,W 4 ,W 5 }. (32)
W 1o min ? -M
?W = o12 ,
?(r - r ) ⊥ ort W .
? 12min r - r w 12
? w 12 (28) The hodograph of potential digging resistive
force HW (Fig. 2b) is a polygon bounded by the
hodographs of boundary digging resistive forces lim-
Boundary digging resistive forces W i (i=3, 4, 5) which can be overcome by the driving mechanisms of the manipulator for a known ort W , and the posi- ited by the excavator stability and the hodographs of boundary digging resistive forces limited by the ma- nipulator driving mechanisms.
934 Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942
The line n-n (Fig. 2b) perpendicular to the cur- rent digging velocity vector v divides the hodograph HW into two parts. The first part is the zone of direc- tions of action, which corresponds to the digging by toward oneself technology and it is called hodograph of effective digging resistances HW e . In this zone, the dot product of the vector of the potential digging resistive forces and the digging velocity vector is negative:
The relative position of the bucket with respect to the stick for rated bucket digging force is deter- mined by angle θ5d .
In the same fashion, the stick digging force F 4d
represents the maximum boundary force which the stick driving mechanism exerts at maximum work pressure of the hydrostatic system. This force is for the direction of the digging force perpendicular to the bucket cut cutting radius in position θ5d when the rated bucket digging force is achieved. The relative W e = v ?W m < 0,
(33)
position of the stick with respect to the boom upon rated stick digging force is determined by angle θ4d . and the dot product of the potential digging forces and the digging velocity vector is positive:
3.4 Corrected digging forces
The corrected digging force F u is defined for F e = v ?W m > 0.
3.3 Rated digging forces
(34)
the entire workspace of the excavator using the fol- lowing equation:
Digging forces rated by the manufacturer are
N 3 N 4
N 5 N w
determined according to standards (ISO 6015, 2006;
F = k ∑∑∑∑k
xy
? k w ? F e srkw
s =1 r =1 k =1 w =1 ,
(35)
SAE J1179, 2008) and they include bucket digging u θ
N ? N N ? N
force F 5d and stick digging force F 4d (Fig. 3).
3
4
5
w
B 4
O 4
p41max
θ4d
F c41max
x 3
where F e srkw is the magnitude of the effective dig- ging force, with subscripts which show that the force relates to a specific position of the boom (s ), stick (r ) and bucket (k ) of the manipulator and a specific di- rection of digging force (w ); k θ is the factor of the digging area in the workspace of the excavator; k w is θ4o
L 4
М
F c51max
the factor of digging force direction, k xy is the factor of position; N 3 is the number of positions of the ma- nipulator boom; N 4 is the number of stick positions; N 5 is the number of bucket positions; and N w is the p51max
R =s
O 5
number of directions of digging resistive force.
F
5d 5 5
F 4d
L 5
O w
θ5d
x 4
x 5
Defined by Eq. (35), the corrected digging force represents the mean value of the effective digging force for the entire workspace of the excavator. It is corrected by the coefficients which takes into ac- count the influence of the following parameters: (1) the size of the excavator workspace; (2) the direction Fig. 3 Rated digging forces of a hydraulic excavator backhoe attachment
The bucket digging force F 5d represents the maximum force which the bucket driving mecha- nism exerts at maximum work pressure of the hydro- static system for the direction of digging force per- pendicular to the bucket cutting radius.
of the digging forces action; and (3) the position of the cutting edge top in the excavator workspace.
The meanings of the correctional factors in Eq. (35) are as follows.
1. Digging area factor k θ
When determining the corrected digging force, one should take into account the relative size of the entire workspace of the excavator by using the
Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942 935
y
digging area factor k θ, which is defined by the ratio:
where θwp =0° and θwk =180° are the minimum and the maximum angles of direction of the digging re- k θ =
θ3o + θ4o + θ5o ,
θ3or + θ4or + θ5or
(36)
sistive force, respectively.
3. Digging position factor k xy
Excavators of all sizes do not work in a uniform where θ3o , θ4o , and θ5o are angles of ranges of rela- tive motion of manipulator boom, stick, and bucket, θ3or , θ4or , and θ5or are referential constant angles of ranges of relative movement of manipulator boom, stick, and bucket selected so that they have slightly larger values than the potential appropriate angles θ3o , θ4o , and θ5o .
2. Factor of direction of the digging force k w
It is known that in the hydraulic excavators which work according to toward oneself technology, the direction of the digging resistive force is variable and has random values, which is usually dispersed around the direction of the vector of the digging ve- locity (Dudczak, 1977). This points out that digging fashion in every part of their workspace for their lifecycle. There are zones where, they usually per- form the toward oneself technology. According to this technology, to determine the corrected digging force, it is necessary to introduce the digging posi- tion factor k xy , whose value is defined as a function of the coordinates x w , y w of the bucket tooth top within the boundaries: 0≤k xy ≤1. For the digging to- ward oneself technology, the digging position factor k xy is determined in relation to the radius R and angle ψ (Fig. 5) of the polar coordinate system with a cen- ter in point O 3.
k Rn
k ψn
forces in all possible directions covered by the hodo- graph of digging forces do not have the same signifi- cance for the toward oneself technology. Therefore, when defining the corrected digging force, it is nec- essary to introduce the factor of direction of the dig-
k R
R )
R n
R i+1
D xy
ging force k w with a value that is defined within the limits 0≤k w ≤1, depending on the digging technology and angle θw , where the digging velocity is taken to be perpendicular to the bucket cutting radius (Fig. 4).
k Ri k R 1
Y
x 3 O 3 R 1
3
O
ψn
R
R i
ψi +1
ψ
ψi
x w
ψ
k ψ(i+1) W k ψ
y w
k
1
ψ1
D xy
2
k ψ2
ψi
k ψ(ψ
Fig. 4 Defining factor of direction of the digging force k w
Fig. 5 Definition of the factor of excavator digging posi- tion k xy
For the digging toward oneself technology, the factor of direction of the digging force k w is defined
by
The value of the digging position factor k xy is determined by
k w = sin θw , ?θw = [θwp ,θwk ],
(37)
k xy = k R ? k ψ ,
(38)
936 Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942
w 3 R
where k R is the digging position factor that depends on the digging position radius R ; and k ψ is the dig- ging position factor that depends on the digging po- sition angle ψ.
The connection between the digging coordi- nates (R , ψ) in the polar and the coordinates in the absolute coordinate system (x w , y w ) of an excavator is determined by (Fig. 5)
R = [(x w - x 3 ) 2 + ( y - y )2 ]0.5
,
(39)
ψ = arctg y w - y
3 ,
x w - x 3
(40)
where x 3 and y 3 are the coordinates of the joint O 3.
By selecting the desired number n R of digging radii R i in the closed interval between the minimum and maximum values (R i =[R 1, R 2, …, R n ]), and by selecting the desired number n ψ of the digging posi- tion angles ψi in the interval of its minimum and maximum values (ψi =[ψ1, ψ2, …, ψn ]), a grid of dig- ging position factor values is formed. These values cover and divide the entire workspace of the excava- tor into a desired number of working zones.
For the selected digging position radii R i and angles ψi , the values of the factor are set within the boundaries: 0≤k Ri ≤1, 0≤k ψi ≤1, so that a digging posi- tion factor k xy is determined for every workspace part of the excavator.
Within the digging zone bounded by adjacent digging position radii R i , R i+1 and adjacent digging position angles ψi , ψi +1, the digging position factors have the values of (Fig. 5):
k = k + (k
- k ) R - R i ,
R
R i
R i +1
R i
i +1 - R i
(41)
?R i +1 ≥ R ≥ R i , i = 1, 2,.., n R -1,
ψ -ψ k = k + (k - k ) i
,
ψ ψ
i
ψ i +1
ψ i ψ i +1 -ψ i
(42)
Fig. 6 Computer program algorithm for analysis of
?ψ i +1 ≥ψ ≥ψ i , i = 1, 2,.., n ψ - 1.
hydraulic excavator digging forces
The boundary digging position radii R 1 and R n and the boundary digging position angles ψ1 and ψn are selected so that the grid of digging position fac- tors covers the entire workspace of the excavator. 4 Analysis To show the possibilities of a corrected digging force for assessment of the excavator digging
Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942 937
Table 1 Parameters of kinematic chain members of computational hydraulic excavator model (Fig. 2)
i L s i t i
m (kg) B (mm) L (mm) i i
e ix e iy e iz s ix (mm) s iy (mm) s iz (mm) t ix (mm) t iy (mm) t iz (mm)
efficiency, an analysis for a hydraulic excavator of following quantities are determined: (1) geometrical mass m =16 945 kg and V =0.6 m 3 bucket capacity is parameters (θ , r r , r ) which define the position of
i i , t i w
conducted and presented. Two variants of the exca- vator with the same parameters of kinematic chain links L i (Table 1) and different parameters of ma- nipulator driving mechanisms C 3j , C 4j , and C 5j (Ta- bles 2–4) were comparatively analysed.
The analysis covered the following variants of excavators: variant A={L , C 31, C 41, C 51} and variant B={L , C 32, C 42, C 52}. The variants of the manipulator driving mechanisms had the identical transformation parameters –hydraulic cylinder piston and connecting rod diameters but different transmission parameters – position coordinates of joints by which hydraulic cylinders and transmission levers were connected to the links.
joint centers and mass centers of the kinematic chain of an excavator; (2) loading moments (M o i , M z i ) and moments (M p i max ) of the driving mechanisms; (3) statistical set of boundary (F g ), potential (F m ) and effective (F e ) digging forces for the entire workspace. The number of values of the statistical set parameters was equal to the product: N w ×N 5×N 4×N 3; (4) cor- rected digging force (F u ) and the standard deviation of the corrected digging force (σF u ) for the entire workspace.
The standard deviation of the corrected digging force σF u is defined with the following equation:
Simulation software
The analysis was performed by a computer pro- σ F u gram, with an algorithm (Fig. 6) that is based on the defined mathematical models of the kinematic chain, driving mechanisms, and digging forces.
(43)
On the basis of input quantities, and by using the cyclic change of numbers N w , N 5, N 4 and N 3, the
The appropriate input files and quantities used
in the analysis are given in Table 5.
L 1
0 0 1 0 1020 0 0 450 0 7500 2000 3200 L 2 0 1 0 410 665 0 ?1360 435 0 6500 L 3 0 0 1 5000 0 0 2260 565 0 1250 L 4 0 0 1 2200 0 0 535 135 0 530 L 5 0 0 1 1300 0 0 580 335 0 455
Note: B is the track gauge, and L is the length between centers of rollers
Table 2 Parameters of variant solutions C 3j of boom drive mechanism of computational excavator model (Fig. 2)
C 3j d 31 (mm) d 32 (mm) c 3p (mm) c 3k (mm) a 3x (mm) a 3y (mm) b 3x (mm) b 3y
(mm)
m c3 (kg) n c3
C 31 115 80 1555 2605 430 ?520 1870 955 205 2 C 32 115 80 1525 2560 545 ?410 1915 990 205 2 Table 3 Parameters of variant solutions C 4j of stick drive mechanism of computational excavator model (Fig. 2)
C 4j
d 41 (mm) d 42 (mm) c 4p (mm) c 4k (mm) a 4x (mm) a 4y (mm) b 4x (mm) b 4y
(mm)
m c4 (kg) n c4 C 41 140 90 1710 2885 ?545 350 1985 1085 280 1 C 42 140 90 1723 2910 ?620 187 **** **** 280 1 Table 4 Parameters of variant solutions C 5j of bucket drive mechanism of computational excavator model (Fig. 2)
C 5j
d 51 (mm) d 52 (mm) c 5p (mm) c 5k (mm) a 5
(mm) b 5x (mm) b 5y (mm) b 45x (mm) b 45y (mm) c 55 (mm) a 55x (mm) a 55y (mm) m c5 (kg) n c5
C 51 115 80 1449 2381 575 1850 470 230 15 515 0 300 190 1 C 52 115 80 1386 2270 510 1770 570 290 30 475 ?62 351 190 1
938 Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942
M p 3m a x a n d M z 3 (k N ·m )
M p 5m a x a n d M z 5 (k N ·m )
M p 4m a x a n d M z 4 (k N ·m )
Table 5 Input quantities of computer program for the partial range (Δθ34, Δθ35) of motion of the ma- nipulator boom. For the relative position of the stick and bucket, determined by angles θ4d , θ5d , the poten- tial digging force F 31 was greater in variant B,
(a)
M p32max
θ3o
Table 1; parameters of variant solutions of manipulator drive mechanisms file are given in Tables 2–4; and radii R i and angles ψi of digging position and their corresponding values of factors k Ri and k ψi are given in Table 6
M z 3
Analysis of the results
The obtained simulation results were analyzed. The analysis shows that the excavator A and excava- tor B have different transmission functions of the driving mechanisms (r c3, r c4, r c5) and different changes in the driving moments of the boom (M p3max ) (Fig. 7a), the stick (M p4max ) (Fig. 7b) and the bucket (M p5max ) (Fig. 7c); whereas they have almost identi- cal ranges of motion of the manipulator links (θ3o , θ4o , θ5o ). The absolute values of maximum driving moments of the boom, stick and bucket mechanisms are greater than the absolute values of the maximum loading moments (M z 3, M z 4, M z 5).
The analysis of the digging forces exerted by
(b)
θ3p
M p31max
θ3 (°)
M p41max
θ4o
M z 4 M p42max
θ3k
the motion of the stick (F 4) (Fig. 8a) and the bucket (F 5) (Fig. 8b), determined under the conditions pre- scribed by the standard (ISO 6015, 2006) show that the excavator A and the excavator B have the same values of rated stick (F 4d ) and bucket (F 5d ) digging forces. They are exerted under different angles of the relative position of the stick in relation to the boom (θ4d ) and the bucket in relation to the stick (θ5d ).
For the excavator variant A the rated stick (F 4d ) (Fig. 9a) and bucket (F 5d ) digging forces could not be exerted in the range (θ3o ) of the manipulator boom
(c) θ4k θ4p
θ4 (°)
M p51max
θ5o
M
since they are greater than the potential digging force F 31=min{F 3,F 1o ,F 1x }. The force F 31 represents the minimum value of the boundary digging force that could be exerted by the driving mechanism of the boom (F 3) and the boundary forces allowed by the z 5
M p52max
θ5k
θ5p
θ5 (°)
excavator stability (F 1o , F 1x ). On the other hand, for Fig. 7 Maximum driving moments M p i max and loading the excavator variant B, the rated stick (F 4d ) and bucket (F 5d ) digging forces could be exerted only in
moments M zi of driving mechanisms of boom (a), stick (b) and bucket (c) of variant A and B excavators
Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942 939
(a)
F 4d
θ4d
θ4p
(a)
F 1
θ4d =?69.3° θ4d =?84.8° θ5d =?13.7° θ5d =?20.2°
F 3
F 31
F 5d
F 4d
F 31
?θ35 ?θ34
(b)
F 5d
θ5d
θ5 F 5 (k N )
F 4 (k N )
F 1, F 3, F 31 (k N )
y w (m )
because digging reaches x w , y w (Fig. 9b) within the range of motion of boom θ3o , and less for variant B in comparison to variant A.
θ4k
θ4d
(b)
θ3 (°)
θ4 (°)
B
A
θ5k
θ5 (°)
θ5d
p
x w (m)
Fig. 9 Rated digging forces
(a) Ratio of rated to potential digging forces in the range of Fig. 8 Rated digging forces of stick F 4d (a) and bucket F 5d
(b) of variant A and B excavators, determined according to standard
The analysis of the hodograph of digging forces for a specified number of the kinematic chain positions in the entire workspace of the excavator was performed for both excavators. For two isolated different manipulator positions defined by general- ized coordinates (θ3, θ4, θ5), the analysis showed that the hodographs of effective digging forces HF e (Fig. 10) were significantly different for variant A and B excavators. The comparative analysis of the spectra of values of effective digging forces F e in the entire working range of variant A (Fig. 11a) and B (Fig. 11b) excava- tors, provides only visual but not quantitative assess- ment of the digging efficiency of the excavators. motion of boom; (b) Trajectory of central bucket cutting edge upon potential action of rated forces in working range of variant A and B excavators
A specific color of the spectrum corresponds to each of the positions of the bucket tooth tip. Each spectrum color represents a certain interval of the effective digging force value. The spectrum shows that in the same position (same coordinates x w , y w ) of the bucket tooth tip, different values of the effective digging force can be achieved, since the same posi- tion of the bucket tooth tip can be achieved by vari-
ous relative positions of boom, stick and bucket. Spectra of effective digging forces are obtained using the Origin Pro 8 program, based on the files of effective digging forces obtained using a developed program whose algorithm is shown in Fig. 6.
Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942 941
y w (m )
F m y (k N )
y w (m )
F m y (k N )
(a)
θ3=?40° θ4=?50° θ5=?10° x w =3.99 m y w =?5.01 m
(b)
θ3=?30° θ4=?70° θ5=?20° x w =3.68 m y w =?4.01 m
F e
HF e
F e
HF e
O w
O w
HF m
HF m
F m x (kN) F m x (kN)
Fig. 10 Hodographs of potential (HF m ) and effective (HF e ) digging forces of variant A and B excavators for two dig- ging positions in workspace
(a)
F e (kN)
(b)
F e (kN)
x w (m)
x w (m)
Fig. 11 Spectrum of values of effective digging forces F e in the entire workspace of excavator
(a) Variant A; (b) Variant B
For the given input conditions and quantities (Table 5), the corrected digging forces F u and the standard deviation of the corrected digging force σF u were determined for variant A and B excavators for constant and variable values of the digging area fac- tor (k θ), the digging position factor (k xy ) and the fac- tor of direction of the digging force (k w ) (Table 6).
In the first case, unit values of correctional fac- tors (k θ=k xy =k w =1) are selected (Table 6), hence, they have no impact on the value of the corrected digging force. For this case, the working range of the
excavator, digging position and the direction of force are neglected, so the corrected digging force, in fact represents the mean value of the effective digging force for the entire workspace of the excavator. For the selected unit values of the correctional factors, the results of the analysis show that the corrected digging force of the excavator variant B is greater, while its standard deviation is smaller than the devia- tion of the excavator variant A.
The variables of the following values are se- lected in the second case: digging area factor (k θ≤1)
940 Janosevic et al. / J Zhejiang Univ-Sci A (Appl Phys & Eng) 2012 13(12):926-942
(Table 6), digging position factor (k xy≤1) and the fac- tor of direction of the digging force (k w≤1), where greater values are given to the digging zones below the level of ground and to the directions of digging force action perpendicular to the bucket cutting ra- dius. The analysis shows that the corrected digging force of the excavator variant A was greater, while its standard deviation was smaller than that of the exca- vator variant B.
Results of the analysis show that when designing an excavator, the defined corrected digging force can be used in the analysis and assessment of possible variant solutions for excavator driving mechanisms.
Factors of the corrected digging force kθ, k xy and k w do not always have the same values but are cho- sen according to the set demands of the excavator digging function with the aim of a more efficient, coordinated and directed action of all driving mechanisms in the desired (priority) zone of the working range of an excavator. When a possible set of variant excavator solutions is analyzed and as- sessed comparatively, the selected values of factors kθ, k xy and k w do not change, where the optimal solu- tion corresponds to the maximum value of the cor- rected digging force.
5 Conclusions
In this paper, quantitative measures for the as- sessment of the hydraulic excavator digging effi- ciency have been introduced.
Rated digging forces of hydraulic excavators, defined according to standards, do not present the full possibilities of the digging function in the entire workspace of an excavator truthfully. They represent the maximum potential digging forces which can be achieved in certain positions of the kinematic chain within the narrow zone of the workspace.
The analysis of the digging forces of hydraulic excavators is somewhat difficult due to the big number of possible positions of the manipulator, dif- ferent digging technologies and various exploitation conditions in the spatial workspace.
Boundary digging forces limited by the excava- tor stability depend on the distribution of the total mass of the excavator and the support contours of the excavator base. The support contours are outlined by the tip over lines of the excavator. Boundary digging forces that can be exerted by the driving mechanisms of the excavator manipulator depend on the transmis- sion functions of the driving mechanisms.
Hodographs of boundary digging forces and hodographs of effective digging forces are suitable for the analysis of digging forces in a specific posi- tion of the excavator kinematic chain. However, hodographs of digging forces cannot be considered as the resulting indicators of digging efficiency for the entire workspace of the excavator.
The corrected digging force defined in this pa- per, represents a contribution to the development of quantitative measures of digging efficiency in the entire workspace of an excavator. In the comparative analysis of the different variants of the identical ex- cavators, the defined corrected digging force show the following: (a) the way that actions of manipulator mechanisms are compatible mutually and in relation to the boundaries conditioned by the excavator sta- bility; (b) how many maximum possibilities of active operations of manipulator mechanisms are corrected toward the desired zones of the working range and expected directions of digging resistive forces.
The results of the analysis obtained by the de- veloped software show that the proposed quantitative measures can be used for assessment of the digging efficiency of the excavators and to serve as an opti- mization criterion for the synthesis of driving mecha- nisms. The defined corrected digging force can be employed in the assessment of the digging efficiency of existing excavator models, and as a criterion dur- ing the synthesis of manipulator driving mechanisms.
Table 6 Corrected digging forces of variant A and B excavators for constant and variable values of factors kθ, k w, k Ri, and kψi
kθk w
k Ri kψi F u (kN) σF u
R i=1.5 m 1.8 m 3.4 m 5.1 m 6.8 m 8.5 m Ψi=?90°?70°?20° 0° 20°90° Variant A Variant B Variant A Variant B
1.0 1.0 1.0 1.0 1.0 1.0 1.0 1.0 1.0 1.0 1.0 1.0 1.0 1.0 69 442 70 689 29 146 29 105
θ3o + θ
4o
+ θ
5o
θ + θ + θ sinθw 0.1 0.2 0.3 1.0 1.0 0.8 0.1 1.0 1.0 0.5 0.2 0.1 23 942 22 001 23 315 24 051 3or 4or 5or
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迄今全球最大的20款液压挖掘机 ——来自【巨车网】论坛:新型号出现后本帖将相应编辑更新,欢迎大家持续关注—— 排序主要依据:工作重量,同时参照功率(有对应升级型号的老版本不再包括) ================================================= 1、CATERPILLAR 6120B H FS MINExpo 2012上,卡特彼勒发布了此型号的超大型液压正铲,创新的混合技术通过使用液电再生技术以及独特的储能系统,预计可使每吨耗油量降低25%。此机型只提供正铲结构,可高效匹配超大型的矿用卡车…… 6120B H FS主要技术数据: 工作重量:1270吨 总功率:3360kW(4500hp) 铲斗容量:46-65立方米 2、CATERPILLAR 6090 FS 2010年3月Bucyrus收购了Terex采矿部,11月时Bucyrus又被CAT以76亿美元价格收购,自此RH400归于卡特旗下并换型为6090 FS。O&K在1997年推出了RH400第一台样机(工作重量为830吨)。6090 FS由两台Cummins QSK60柴油机(2-stage,16缸,涡轮增压)提供动力,油料箱容
准履带板宽度为2米,最大移动速度2.2km/h;回转速度4.0rpm。6090 FS可选电力驱动形式(供电电压6600V,输出功率3200kW)。此型号只提供正铲结构。 6090 FS主要技术数据: * 工作重量:980吨 * 最大功率:3360kW(4570马力)/1800rpm * 铲斗容量:50立方米(载荷84吨) 3、HITACHI EX8000-6 它是Hitachi目前最大的液压挖掘机,首台EX8000于2004年3月开发成功。EX8000-6的操作重量已经突破了800吨,它装有两台16缸、排量60升的Cummins QSK60C柴油引擎,油箱容量为14900升。履带宽度1.85米,最大行走速度2.0km/h,最大挖掘高度20.5米,回转速度3.2rpm,操纵杆轻触式电控形式。此型号目前仅有正铲机型。 EX8000-6主要技术数据: * 工作重量:811吨 * 最大功率:2×1450kW(3888hp)/1800rpm * 铲斗容量:40立方米(载荷62.5吨)
国外液压挖掘机目前水平及发展趋势 工程机械液压网 2011年02月09日评论? 41 views 字体: 小中大 工业发达国家的挖掘机生产较早,法国、德国、美国、俄罗斯、日本是斗容量3.5-40m3单斗液压挖掘机的主要生产国,从20世纪80年代开始生产特大型挖掘机。例如,美国马利昂公司生产的斗容量50-150m3剥离用挖掘机,斗容量132m 3的步行式拉铲挖掘机;B-E(布比赛路斯-伊利)公司生产的斗容量168.2m3的步行式拉铲挖掘机,斗容量107m3的剥离用挖掘机等,是世界上目前最大的挖掘机。 从20世纪后期开始,国际上挖掘机的生产向大型化、微型化、多功能化、专用化和自动化的方向发展。 1)开发多品种、多功能、高质量及高效率的挖掘机。为满足市政建设和农田建设的需要,国外发展了斗容量在0.25m3以下的微型挖掘机,最小的斗容量仅在 0.01m3。另外,数量最的的中、小型挖掘机趋向于一机多能,配备了多种工作装置——除正铲、反铲外,还配备了起重、抓斗、平坡斗、装载斗、耙齿、破碎锥、麻花钻、电磁吸盘、振捣器、推土板、冲击铲、集装叉、高空作业架、铰盘及拉铲等,以满足各种施工的需要。与此同时,发展专门用途的特种挖掘机,如低比压、低嗓声、水下专用和水陆两用挖掘机等。 2)迅速发展全液压挖掘机,不断改进和革新控制方式,使挖掘机由简单的杠杆操纵发展到液压操纵、气压操纵、液压伺服操纵和电气控制、无线电遥控、电子计算机综合程序控制。在危险地区或水下作业采用无线电操纵,利用电子计算机控制接收器和激光导向相结合,实现了挖掘机作业操纵的完全自动化。所有这一切,挖掘机的全液压化为其奠定了基础和创造了良好的前提。 3)重视采用新技术、新工艺、新结构,加快标准化、系列化、通用化发展速度。例如,德国阿特拉斯公司生产的挖掘机装有新型的发动机转速调节装置,使挖掘机按最适合其作业要求的速度来工作;美国林肯贝尔特公司新C系列LS-5800 型液压挖掘机安装了全自动控制液压系统,可自动调节流量,避免了驱动功率的浪费。还安装了CAPS(计算机辅助功率系统),提高挖掘机的作业功率,更好地发挥液压系统的功能;日本住友公司生产的FJ系列五种新型号挖掘机配有与液压回路连接的计算机辅助功率控制系统,利用精控模式选择系统,减少燃油、发动机功率和液压功率的消耗,并处长了零部件的使用寿命;德国奥加凯(O&K)公司生产的挖掘机的油泵调节系统具有合流特性,使油泵具有最大的工作效率;日本神钢公司在新型的904、905、907、909型液压挖掘机上采用智能型控制系统,即使无经验的驾驶员也能进行复杂的作业操作;德国利勃海尔公司开发了ECO(电子控制作业)的操纵装置,可根据作业要求调节挖掘机的作业性能,取得了高效率、低油耗的效果;美国卡特匹勒公司在新型B系统挖掘机上采用最新的3114T型柴油机以及扭矩载荷传感压力系统、功率方式选择器等,进一步提高了挖掘机的作业效率和稳定性。
世界上最大液压挖掘机有多重
世界上最大液压挖掘机有多重 液压挖掘机是一种多功能机械,它在减轻繁重的体力劳动,保证工程质量,加快建设速度以及提高劳动生产率方面起着十分重要的作用。那世界上液压挖掘机到底有多重呢?下面由我来告诉你答案吧。 ? RH400是目前市场上最大的液压挖掘机,是TEREX/O&K公司的旗舰产品,也是当今唯一的1000吨级液压挖掘机。RH400操作重量980吨,发动机输出功率达3280千瓦,正铲能容纳85吨的矿物,绝对是个庞然大物!它在加拿大油砂矿创造了液压正铲新的世界纪录:在性能测试时大大超过9000吨/小时,平均产量超过5500吨/小时,对于超大型的卡车例如CAT 797B等,3-5铲即可装满。RH400挖掘机可配备柴油发动机或电机驱动装置。 液压挖掘机六大系统 传动系统 单斗液压挖掘机在建筑、交通运输、水利施工、露天采矿及现代化军事工程中都有十分广泛的应用,是各种土石方施工中不可缺少的主要机械
设备。 流体传动形式 流体传动包括以下三种形式: 1、液压传动—借助于液体的压力能来传递动力和运动的传动形式; 2、液力传动—借助于液体的动能来传递动力和运动的传动形式;如液力变矩器 3、气压传动—借助于气体的压力能来传递动力和运动的传动形式。 传动技术优点 液压传动是以液压油为工作介质进行能量传递和控制的一种传动形式,通过各种液压元件组成不同功能的基本回路,再由这些基本回路组成能够满足各种要求的液压系统。 液压传动与机械传动和电力拖动系统比,其主要优点有: 1、元件布局安装有很大的灵活性,能构成复杂系统; 2、可实现大范围无级调速,范围可达2000:1;
Unit7 Hydraulic Excavators 液压挖掘机 7.1Overview概述 7.1.1Basic Concept基本概念 An excavator is an engineering vehicle consisting of an articulated arm(boom,stick),bucket and cab mounted on a pivot(a rotating platform)atop an undercarriage with tracks or wheels. Their design is a natural progression from the steam shovel. 挖掘机是一种由铰接臂杆(动臂和斗杆)、铲斗和安装于履带或轮式底盘上的转盘(一种旋转平台)所组成的工程机械(车辆)。挖掘机是在蒸汽铲的基础上自然发展起来的。 The history of heavy excavating machinery began in1835when the dipper shovel was invented to excavate hard soil and rock and to load trucks.Of course,with the invention of gasoline-and diesel-powered vehicles,construction equipment became even more adaptable.Most construction equipment is powered by diesel engines,although electric-power,battery power,and propane tanks are used on specialized equipment. 重型挖掘机的历史始于1835年,当时发明了拉铲式挖掘机用于开挖坚硬的土石方及装载卡车。当然,随着汽油机和柴油机车辆的发明,工程机械也变得越来越适用。虽然在一些专用设备上使用了电力驱动、蓄电池驱动和丙烷气罐,然而大多数工程机械仍然依靠柴油机驱动。 Design modifications are driven by customer demand.As of2000,the two primary areas where customers would like to see more improvements are in the ease of operation and the operator's comfort.The need for simple operation is forced by the fact that there are fewer skilled operators in the marketplace.And operations and reliability are both improving because of the
摘要 参照国内外著名品牌的先进机型,收集分析挖掘机的尺寸参数,性能参数,作业要求等相关资料。以此为基础,进行对挖掘机的机械结构,液压系统,电气控制的设计等。设计思路是从挖掘机总体的工作性能和动作要求入手,确定整机必要的性能参数并对整机的稳定性进行校核。随后将对挖掘机工作装置的具体结构、基于单片机对柴油机油门控制系统进行设计并运用软件将其绘制。 关键词:液压挖掘机,总体设计,单片机,柴油机油门控制
ABSTRACT According to the domestic and international famous brands of advanced type,collecting and analyzing the information of the hydraulic excavator about their size, performance parameters, operational requirements other relevant information. On this basis, I designed the mechanical structure, hydraulic excavator, electrical control system of hydraulic excavator and so on. The way to solve the task is on the basis of the function of the whole of the excavator ,calculated the performance parameters of the excavator and checked the stability of the excavator. Then I designed the structure of the platform of type excavator, desgin control system based on MCU(Micro Controller Unit),and use software to draw. Key words:Hydraulic excavator, overall design, micro-controller unit (MCU), control of diesel engine’s accelerator
70吨大型液压挖掘机动臂有限元分析 一、动臂计算工况 挖掘机在工作过程中,作业对象千变万化,土质及施工现场也各异,其工作装置运动与受力情况比较复杂。故选择了最危险工况来进行强度校核。 工况一: 1)、动臂位于最低(动臂油缸全缩); 2)、斗齿尖、铲斗与斗杆铰点、斗杆与动臂铰点三点位。 图1 工作装置挖掘姿态(工况一、二) 工况二: 在工况一的基础上:3)斗边点遭遇障碍,侧向力W k。 工况三: 1)、动臂位于动臂液压缸作用力臂最大处; 2)、斗杆油缸作用力臂最大(斗杆油缸与斗杆尾部夹角为90°); 3)铲斗发挥最大挖掘力位置,进行正常挖掘。 工况四:
在工况三的基础上:3)斗边点遭遇障碍,侧向力W k。 图2 工作装置挖掘姿态(工况三) 三、斗杆受力分析 1)、斗杆铰点载荷的确定 ①计算工况一:θ1=-41.68°,θ2=131.684°,θ3=180°。 从重庆大学反铲分析软件中提取进行有限元分析所需要的数据:动臂缸作用力为:-927.87kN; 斗杆缸作用力为:423.386KN; 铲斗缸作用力为:318.225kN; 动臂油缸铰点:Rx=915.301 kN;Ry=150.391 kN; 斗杆油缸铰点:Rx=264.623 kN;Ry=330.5 kN; 斗杆动臂铰点:Rx=-370.888 kN;Ry=-376.969 kN;
图3挖掘工况一、二示意图 ②计算工况二: 在工况一的基础上,加上侧齿障碍产生的弯矩和扭矩,及侧向力W k=23.638 KN。 ③计算工况三:θ1=4.504°,θ2=111.993°,θ3=151.371°。 图4 挖掘工况三示意图 从重庆大学反铲分析软件中提取进行有限元分析所需要的数据:动臂缸作用力为:-927.87 kN;
□ 世界液压挖掘机发展 纵不雅世界工业化汗青,现代挖掘机械的源头可以追溯到15百年末年,意大利锡耶纳(距佛罗伦萨南部约50公里)出现的淤沙挖掘船由于水上航运一直是意大利非常重要的交通体式格局,这就需要经常疏浚河流按照《达·芬奇条记》记载,这类挖掘船由两艘小舟平列,侧舷之间架设水车是的转轮,转轮四角伸出的木杆端部安装铲斗,以人力驱动转轮旋转,让铲斗探入河底铲挖淤沙,铲斗容量一般不跨越0.2-0.3立米米当扑满淤沙的铲斗旋转至顶端时,淤沙会经由过程滑板全部倒出进船里面中这应该是多斗式挖掘机的最新大概的形状1712年英国人托马斯·纽科门(Thomas Newcomen)发现活塞式蒸汽轮机,后经瓦特改进,于1769年制成瓦特式蒸汽轮机,其能效比纽科门蒸汽轮机提高了5倍以上蒸汽轮机的广泛应用,使挖掘机械起头由人力驱动向机械驱动转变 1833-1835年代里,美国费城的铁路工程师威廉?奥蒂斯(William Otis),设计和打造了熬头台以蒸汽轮机驱动、安装在铁路平车上的吊臂式单斗挖掘机其接纳铁木混合结构,吊臂回转依然靠人力用绳牵引,经由过程不断延伸铁轨实现带状开挖,是以被称为为铁路铲(蒸汽铲)但由于奥蒂斯英年早逝,和专利保护、人力成本低廉等因素,奥蒂斯式蒸汽铲没有推广应用直至1870年往后,美国大规模建设铁路,蒸汽铲的发展步入黄金时代,性能获得精益求精,起头应用于铁路建设、开挖运河、露天矿剥离等领域1910年美国出现了熬头台电机驱动的蒸汽铲,并起头应用履带行走装置1912年出现了燃料机和煤油机驱动的全回转式蒸汽铲,1916年出现了柴油发电机驱动的蒸汽铲1924年柴油机直接驱动起头用于单斗挖掘机上此后,随着汽车工业的发展,轮胎式底盘起头慢慢应用于小规模挖掘机上20百年40年代出现了在拖拉机上配装液压反铲的吊挂式挖掘机 1835年先后,美国人威廉?奥蒂斯研究制造的蒸汽挖掘机(蒸汽铲),成为现代挖掘机的鼻祖 1951年,法国Poclain(波克兰)公司推出世界熬头台全液压正铲挖掘机 第二次世界大战和战后恢复建设,有力鞭策了挖掘机械发展随着液压技能的应用,1951年,法国Poclain(波克兰)公司推出世界熬头台全液压挖掘机波克兰公司首创人——乔治?巴塔伊(Georges Bataille)生于1879年,他于1927年,和工程师安东尼?莱杰合股成立公司,主要营业为修理农用机械,1930年公司改名为波克兰打造公司波克兰从1948年起头打造小规模轮式挖掘机,员工数量增至120人1951年10月,波克兰研究制造熬头台正铲液压挖掘机,它接纳道奇4x4改装的轮式底盘,前部为汽车驾驶室,后部为液压挖掘机由节制台、液压臂、铲斗等构成的上车体,盘绕底盘后轴上方的立柱旋转,正方铲斗容积约1立米米 1961年波克兰公司推出TY45型轮式液压挖掘机,接纳奇特的前三轮式底盘,总功率48Hp(马力),重量10吨,该机型至1982年共销售了3万台1974年波克兰公司的挖掘机营业被美国凯斯集团吞并,仅保留了液压元件营业凯斯创立于1842年,是世界熬头家生产蒸汽轮机式脱粒机的企业,1912年起头生产蒸汽压路机、平地机等产物1957年凯斯收购了印第安那州的美国拖拉机公司,当时这家公司已经开发出了履带式液压挖掘机 初期试制的液压挖掘机多接纳飞机和机床的液压技能,缺少适用于挖掘机各种工况的液压元件,打造质量不够不变,配套件也不齐备从20百年60年代起,液压挖掘机步入推广和热潮阶段,各国挖掘机打造厂和品种增加很快,产量猛增到70年代初,液压式挖掘机已占挖掘机总产量的83%,慢慢取代了机械式挖掘
大型液压挖掘机海外市场研究与分析 大约从上世纪70年代开始,日本日立建机公司开始潜心研究大型液压挖掘机,并且一直以来都取得了很大的突破。经过30多年的发展,目前超大型液压挖掘机的最大工作重量已突破900t级,铲斗斗容达50m3。由于具有结构紧凑、操作方便、运动灵活及易于维护保养等优点,超大型液压挖掘机已形成逐步代替钢索机械式或电动式挖掘机(俗称电铲)的趋势。超大型液压挖掘机主要用于各种大规模露天矿山的开采及大型基础建设,同时还被用于填海造地工程及港湾河道疏通工程,其中正铲式挖掘机占大部分。目前生产超大型挖掘机的厂商主要分布在美国、日本、德国等工程机械制造强国,著名的生产企业有利勃海尔、卡特彼勒、日立、小松、特雷克斯等公司。1987年,当时世界最大的日立EX3500型正铲超大型挖掘机(斗容量18.8m3,工作重量330t)在日本开发成功,并在这一年又开发了斗容量22m3,工作重量420t的EX4200型正铲超大型挖掘机。此时世界其他公司也相继开发生产出超大型液压挖掘机,如小松PC1600型、卡特彼勒5130型,利勃海尔R992型、R994型,特雷克斯RH20C型、RH90C型等。1990年以后,超大型液压挖掘机的斗容量不断增加,达到了50m3以上;机重不断加大,达800t以上,目前已经形成超大型矿山型与土方型液压挖掘机系列。日立建机于2004年成功研发面向矿山用户的EX8000超大型液压挖掘机,铲斗容量为40m3,大约相当于6辆11吨翻斗车的容量,最大工作重量780t,是目前日本制造的最大级别的液压挖掘机。该超大型液压挖掘机不仅可顺应为降低采掘运输成本而不断加大矿山自卸车载荷的趋势,而且与钢索式电动挖掘机相比,超大型液压挖掘机具有更好的机动机和操作性,在降低运营成本方面也极具优势。至今,日立建机公司已经向北美、澳大利亚、亚洲和南非等大型矿山交付了大约700台多种型号的超大型液压挖掘机,在该级别中占据了全球最大的市场份额。超大型液压挖掘机的快速发展与应用主要得益于其具有良好的工作性能和耐久性、可靠性,极大地提高了挖掘机工作效率和经济效益。优越的操纵控制性及作业性能是超大型液压挖掘机的显著特点。超大型液压挖掘机采用发动机-油泵电气控制系统,可以最大限度地发挥发动机的性能、降低油耗,使机械高效地工作。大部分超大型挖掘机采用故障检测系统,利用装设在驾驶室的显示器为驾驶员提供各主要部件、系统及整机的工作状况,机械操作的自动化程度及适应性大大提高。另外,超大型液压挖掘机的工作装置也有了极大的改进,在斗杆与铲斗之间采用了平行伸缩机构,使铲斗的工作范围加大,作业性能提高。为使驾驶员容易观察到大型自卸车的装卸状况,许多超大型挖掘机驾驶室的前窗采用了前倾的形式,以使之具有良好的观察视野。为改善驾驶员的工作条件,驾驶室内均采用了大容量空气调节器,使驾驶员可以长时间工作而不致感到疲劳。在一些超大型挖掘机上采用履带接地长度可伸缩机构,按需改变挖掘机的接地面积,大大提高挖掘机的作业性能。近年来,国内外大型矿山及水利工程建设用的非公路自卸汽车在不断向大型化方向发展,目前使用的多是120t-220t级及更大吨位的自卸汽车。为了适应自卸汽车大型化的趋势,保证在联合作业中与自卸汽车相匹配,超大型挖掘机势必提高斗容量,使单位时间内的工作量增加,从而有较高的生产效率。对于超大型液压挖掘机,良好的维修及保养性与提高生产率、降低成本有着直接的关系,也是其重要的性能之一。目前,超大型挖掘机根据机械结构及部件均设立多个监测项目,驾驶员可以通过设置
国外液压挖掘机目前水平及发展趋势 工业发达国家的挖掘机生产较早,法国、德国、美国、俄罗斯、日本是斗容量3.5-40m3单斗液压挖掘机的主要生产国,从20世纪80年代开始生产特大型挖掘机。例如,美国马利昂公司生产的斗容量50-150m3剥离用挖掘机,斗容量132m 3的步行式拉铲挖掘机;B-E(布比赛路斯-伊利)公司生产的斗容量168.2m3的步行式拉铲挖掘机,斗容量107m3的剥离用挖掘机等,是世界上目前最大的挖掘机。 从20世纪后期开始,国际上挖掘机的生产向大型化、微型化、多功能化、专用化和自动化的方向发展。 1)开发多品种、多功能、高质量及高效率的挖掘机。为满足市政建设和农田建设的需要,国外发展了斗容量在0.25m3以下的微型挖掘机,最小的斗容量仅在 0.01m3。另外,数量最的的中、小型挖掘机趋向于一机多能,配备了多种工作装置——除正铲、反铲外,还配备了起重、抓斗、平坡斗、装载斗、耙齿、破碎锥、麻花钻、电磁吸盘、振捣器、推土板、冲击铲、集装叉、高空作业架、铰盘及拉铲等,以满足各种施工的需要。与此同时,发展专门用途的特种挖掘机,如低比压、低嗓声、水下专用和水陆两用挖掘机等。 2)迅速发展全液压挖掘机,不断改进和革新控制方式,使挖掘机由简单的杠杆操纵发展到液压操纵、气压操纵、液压伺服操纵和电气控制、无线电遥控、电子计算机综合程序控制。在危险地区或水下作业采用无线电操纵,利用电子计算机控制接收器和激光导向相结合,实现了挖掘机作业操纵的完全自动化。所有这一切,挖掘机的全液压化为其奠定了基础和创造了良好的前提。 3)重视采用新技术、新工艺、新结构,加快标准化、系列化、通用化发展速度。例如,德国阿特拉斯公司生产的挖掘机装有新型的发动机转速调节装置,使挖掘机按最适合其作业要求的速度来工作;美国林肯贝尔特公司新C系列LS-5800 型液压挖掘机安装了全自动控制液压系统,可自动调节流量,避免了驱动功率的浪费。还安装了CAPS(计算机辅助功率系统),提高挖掘机的作业功率,更好地发挥液压系统的功能;日本住友公司生产的FJ系列五种新型号挖掘机配有与液压回路连接的计算机辅助功率控制系统,利用精控模式选择系统,减少燃油、发动机功率和液压功率的消耗,并处长了零部件的使用寿命;德国奥加凯(O&K)公司生产的挖掘机的油泵调节系统具有合流特性,使油泵具有最大的工作效率;日本神钢公司在新型的904、905、907、909型液压挖掘机上采用智能型控制系统,即使无经验的驾驶员也能进行复杂的作业操作;德国利勃海尔公司开发了ECO(电子控制作业)的操纵装置,可根据作业要求调节挖掘机的作业性能,取
液压挖掘机市场分析研究 发表时间:2018-11-02T16:34:36.923Z 来源:《知识-力量》2018年11月下作者:梁磊[导读] 液压挖掘机由各种设备组成,包括底盘、铲斗和臂动等,它是通过车轮或轨道实现远距离移动的。液压挖掘机主要用于基础挖掘、钻井、分级等施工活动。在不久的将来,德国、英国和法国等欧洲国家有望在全球液压挖掘机市场表现出稳步增长。中东和非洲的市场随着对设备需求的增加,有利于增长,其次是土耳其,其次是沙特阿拉伯和阿联酋。关键词: (太原重工股份有限公司,山西太原 030024) 摘要:液压挖掘机由各种设备组成,包括底盘、铲斗和臂动等,它是通过车轮或轨道实现远距离移动的。液压挖掘机主要用于基础挖掘、钻井、分级等施工活动。在不久的将来,德国、英国和法国等欧洲国家有望在全球液压挖掘机市场表现出稳步增长。中东和非洲的市场随着对设备需求的增加,有利于增长,其次是土耳其,其次是沙特阿拉伯和阿联酋。关键词:液压挖掘机;市场;需求 液压挖掘机是一种用于挖掘和拆除目的的大型车辆。此外,液压挖掘机可有效地运输大量矿物,将矿石和材料从一个地方到另一个地方,液压挖掘机也被称为挖掘机。液压挖掘机由各种设备组成,包括底盘、铲斗和臂动等,它是通过车轮或轨道实现远距离移动的。液压挖掘机主要用于基础挖掘、钻井、分级等施工活动。所有的运动和功能液压挖掘机由液压缸、液压缸和液压马达组成。液压挖掘机可应用于采矿、基础设施和其他类似部门。现代液压挖掘机有各种各样的尺寸。 1 液压挖掘机现状 目前,超大型挖掘机为了获得最大装载能力,大部分采用的是大容量正铲斗。随着液压挖掘机技术的发展,超大型反铲液压挖掘机开发与市场需求增加引起了行业的关注。由于液压挖掘机具有结构紧凑、操作方便、运动灵活及易于维护保养等优点,超大型液压挖掘机已形成逐步代替钢索机械式或电动式挖掘机的趋势超大型液压挖掘机主要用于各种大规模露天矿山的开采及大型基础建设,同时还被用于填海造地工程及港湾河道疏通工程,其中正铲式挖掘机占大部分。目前生产超大型挖掘机的厂商主要分布在美国、日本、德国等工程机械制造强国。 2 液压挖掘机的优势 超大型液压挖掘机的快速发展与应用主要得益于其具有良好的工作性能和耐久性、可靠性,极大地进步了挖掘机工作效率和经济效益。优越的操纵控制性及作业性能是超大型液压挖掘机的明显特点。超大型液压挖掘机采用发动机-油泵电气控制系统,可以最大限度地发挥发动机的性能、降低油耗,使机械高效地工作。大部分超大型挖掘机采用故障检测系统,利用装设在驾驶室的显示器为驾驶员提供各主要部件、系统及整机的工作状况,机械操纵的自动化程度及适应性大大进步。超大型液压挖掘机的工作装置也有了极大的改进,在斗杆与铲斗之间采用了平行伸缩机构,使铲斗的工作范围加大,作业性能进步。为使驾驶员轻易观察到大型自卸车的装卸状况,很多超大型挖掘机驾驶室的前窗采用了前倾的形式,以使之具有良好的观察视野。为改善驾驶员的工作条件,驾驶室内均采用了大容量空气调节器,使驾驶员可以长时间工作而不致感到疲惫。在一些超大型挖掘机上采用履带接地长度可伸缩机构,按需改变挖掘机的接地面积,大大进步挖掘机的作业性能。近年来,国内外大型矿山及水利工程建设用的非公路自卸汽车在不断向大型化方向发展,目前使用的多是120t-220t级及更大吨位的自卸汽车。为了适应自卸汽车大型化的趋势,保证在联合作业中与自卸汽车相匹配,超大型挖掘机势必进步斗容量,使单位时间内的工作量增加,从而有较高的生产效率。 对于超大型液压挖掘机,良好的维修及保养性与进步生产率、降低本钱有着直接的关系,也是其重要的性能之一。目前,超大型挖掘机根据机械结构及部件均设立多个监测项目,驾驶员可以通过设置在驾驶室里的显示器了解各部分的状态,如机械结构、液压系统、行走及回转部分及铲斗等工作状况。为了易于进行日常维修保养,超大型挖掘机装有很多自动润滑装置,如某液压挖掘机设有31个检测项目、55个自动润滑装置,5000L的燃料箱分成了13个部分,利用集中给排油系统向油箱内注油,保证燃油的正常供给。 3液压挖掘机市场 全球液压挖掘机市场可以细分为基础设施市场和采掘市场,全球液压挖掘机市场的动态增长购买力已带来迅速和可观的投资回报。由于中国和印度等发展中国家基础设施投资不断增加,对液压挖掘机的需求正以非常强劲的速度增长,预计在不久的将来也将继续,上述因素预计将带动全球液压挖掘机市场。 另一方面,建筑设备行业也面临着诸多挑战:成本、经济放缓和严格的政府和环境排放规范,这些因素有可能抑制全球液压挖掘机市场的发展。全液压挖掘机市场以亚太地区为主,其次是北美和欧洲。由于不断增加的基础设施投资和发展中的城市化印度和中国等国对液压挖掘机的需求强劲增长。这反过来又会刺激亚太地区对液压挖掘机的需求。在不久的将来,德国、英国和法国等欧洲国家有望在全球液压挖掘机市场表现出稳步增长,中东和非洲的市场随着对设备需求的增加,有利于增长,其次是土耳其,其次是沙特阿拉伯和阿联酋。在全球液压挖掘机市场的市场参与者包括卡特彼勒公司、迪尔公司和日立工程机械有限公司、三一集团、小松美国公司、徐工集团、太原重工股份有限公司。 4 总结 近年来,随着液压挖掘机技术的发展,超大型反铲液压挖掘机开发与市场需求增加引起了行业的关注,超大型挖掘机为了获得最大装载能力,大部分采用的是大容量正铲斗,与正铲挖掘机相比,反铲挖掘机由于反铲挖掘力稳定,可进行强有力的挖掘,因此具有较高的生产效率。如某型号超大型反铲挖掘机,其斗容量为40.3m3,挖掘高度可达15.2m。 我国超大型液压挖掘机的研发起步较晚,目前国产大型及超大型液压挖掘机超过60t级的极少。太原重工股份有限公司是我国开发大型和超大型液压挖掘机的企业之一,近几年,该公司相继成功开发出拥有自主知识产权的WYD/WY260、WYD/WY390、WYD/WY600、WYD/WY800等系列液压挖掘机。液压挖掘机现在和未来都将作为主要的挖掘设备,我们有必要在此领域投入更多的研究,掌握核心关键技术,占领液压挖掘机更多市场份额。 参考文献 [1]黄昌松.我国液压挖掘机技术引进史实(下)[J].交通世界(建养.机械),2014(Z1):58-61.
第一章绪论§1.1 概述 产品设计是产品生产的第一道工序。传统的产品的生产过程是:首先由设计者根据个人经验初步设计出产品、或者在已有的产品基础上进行模仿、或者改进已有的产品设计出新产品,然后做出模型或样品,再进行试验,对设计上的问题进行改进,重新设计、制造、试验和分析,不但要耗费大量的时间,也要耗费大量的人力和物力。在设计中要求对机器的工作原理、功能、结构、零部件设计、甚至加工制造和装配都要确定下来。虽然不同的设计者可能有不同的设计方法和设计步骤,机械设计的共性规律是客观存在的,其一般步骤是:目标预测、方案设计、技术设计、加工设计、试生产。它需要不断地总结和完善。用传统的设计方法,产品设计质量和风格在很大程度上受设计人员水平的局限,有时严重限制设计质量的提高。 单斗挖掘机是一种重要的工程机械,广泛应用在房屋建筑、道路工程、水利建设、农田开发、港口建设、国防工事等的土石方施工和矿山采掘工业中,对减轻繁重的体力劳动、保证工程质量、加快建设速度、提高劳动生产率起巨大作用。单斗挖掘机分机械传动和液压传动两种。机械传动挖掘机已有一百多年历史,近一、二十年来,随着液压传动技术在工程机械上的广泛应用,单斗液压挖掘机有了通速发展,在中小型单斗挖掘机中,液压挖掘机几乎取代了机械传动挖掘机,大型单斗液压挖掘机也应用日广,这是由于液压挖掘机具有重量轻、体积小、结构紧凑、挖掘力大、传动平稳、操纵简便,以及容易实现无级变速和自动控制等一系列优点。 挖掘机要完成其独特的功能,大部分零件结构复杂,工作条件恶劣,这些零件的结构分析和设计是一件比较困难的工作。全回转步履式液压挖掘机( 以下简称步挖机) 为斗容量小于0. 6m3 的小型挖掘机, 与一般挖掘机的区别在于下车没有行走和转向机构,采用4 个支脚支承整机, 依靠工作装置和回转机构的联动实现机械的前进、后退和转向。它主要由工作装置、平台、回转机构、下车、动力装置和液压系统等组成。其主要特点是结构简单,质量轻,故障少,性能良好, 成本只有同等级履带式挖掘机的 50 %60 %,便于制造和维护;缺点是行走和转向速度慢。步挖机在结构形式和参数选择时应考虑以下要求: (1)整机具有较高的稳定性, 在全域内(360°) 挖掘性能良好; (2)步履行走性好, 即步挖机能开进没有道路的施工现场; (3)支脚调整简便、迅速、适应性强, 安全可靠; (4)小型步挖机长距离转场移动时, 能自行上、下运输车辆; (5)为了防止支脚沉陷和挖掘时的水平移动, 支承爪上部为水平板, 下部为放射状的 爪; (6)能在靠近建筑物的边角和狭小场地上挖掘作业。 (7)其工作简图如图1-1 所示:
单斗液压挖掘机发展概况 挖掘机械是以开挖土方、石方为主要用途的一种通用的或专用的机械设备。是国民经济建设中最广泛使用的重要工程机械。应用于矿山开采、水利水电施工、工业与民用建筑、交通运输工程、油田建设、港口建设、农田土壤改良、林业工程、国防建设等工程中。对于节省人力、减轻体力劳动、保证工程质量、缩短工期、降低工程造价和提高劳动生产率,具有非常重要的作用。 挖掘机械有两大类型:间歇作业式(单斗)挖掘机和连续作业式(多斗)挖掘机。间歇作业的单斗挖掘机有机械传动的单斗挖掘机(以下简称机械式挖掘机)和液压传动的单斗挖掘机(以下简称液压挖掘机)两大类。连续作业式挖掘机包括斗轮式挖掘机、横向或纵向链斗式挖掘机、轮斗式挖沟机、滚切式挖掘机和隧道联合掘进机等。工程施工中应用最广泛的是间歇作业式单斗挖掘机,约占挖掘机械总量的90%以上。目前,单斗挖掘机中应用最广泛的是液压挖掘机,约占单斗挖掘机平均总产量的98%以上。 1.国外液压挖掘机的发展简况 自1836年蒸气驱动的机械式单斗挖掘机问世以来,至今由动力驱动的机械式单斗挖掘机已有170多年的历史。而液压挖掘机的发展却只有60多年的历史。 1951年法国poclain公司首先生产了全回转的、部分机构由液压驱动的(半液压)挖掘机。1954年德国Demag公司制造出世界
上第一台全液压B—504型挖掘机,机重15t,功率约50 ph,斗容量0.4 m3,挖掘机的所有工作机构全部由液压驱动。由于当时液压传动技术尚处于探索阶段,适合挖掘机用的液压元件还不过关,因此影响了液压挖掘机的发展。1963年在西欧市场上,液压挖掘机仅占挖掘机的总产量15%左右。到了上世纪六十年代中、后期,由于解决了液压技术上的问题,促使中、小型液压挖掘机迅速发展,德国、美国、日本、法国等许多国家都在生产液压挖掘机。1968—1970年间液压挖掘机的产量已占挖掘机总产量的88%以上。上世纪七十年代以后,中、小型液压挖掘机的结构经过不断试验和改进更加趋于成熟;研制出高压力、耐冲击负荷的液压元件和液压系统,以及可靠耐用的传动零部件,改进挖掘机的生产制造技术和工艺,大大地提高了液压挖掘机的使用寿命和可靠性;许多国家和公司实现了液压挖掘机的系列化、标准化和通用化;一些国家相继研制出机重100 t以上的大型液压挖掘机,如1979年德国O+k公司的RH300型挖掘机(机重475 t,功率1735 kw,正铲斗容22 m3,挖煤斗容34 m3)。 上世纪八十年代以后到现在的三十年间,液压挖掘机从技术设计到生产工艺和实验研究、从整机到零部件结构、从内部系统的结构和工作性能到机体外观、从使用性能参数到使用可靠性和寿命、从制造厂商数的增长到产品销售形式和售后服务,都发生了重大的、空前的变化,是一次真正的液压挖掘机技术革命。液压挖掘机成为工程中的智能机械,工程机械的领头兵,被广泛应用在各种各样的工程领域。挖掘土石方的工程中大约55%~65%的作业量靠液压挖掘机来完
国内外超大型液压挖掘机发展概况及展望 导读:国外大型液压挖掘机的开发大约从上世纪70年代开始。如日本日立建机公司在1972-1976年间开发的UH12型正铲大型液压挖掘机,其斗容量为2.2m3,工作重量36t;UH20型斗容量3.2m3,工作重量50t;UH30型斗容量4.4m3,工作重量75t。1979年, 国外大型液压挖掘机的开发大约从上世纪70年代开始。如日本日立建机公司在1972-1976年间开发的UH12型正铲大型液压挖掘机,其斗容量为2.2m3,工作重量36t;UH20型斗容量3.2m3,工作重量50t;UH30型斗容量 4.4m3,工作重量75t。1979年,日立建机成功开发了UH50型正铲超大型液压挖掘机,斗容量达8.2m3,工作重量175t。经过30多年的发展,目前超大型液压挖掘机的最大工作重量已突破900t级,铲斗斗容达50m3。由于具有结构紧凑、操作方便、运动灵活及易于维护保养等优点,超大型液压挖掘机已形成逐步代替钢索机械式或电动式挖掘机(俗称电铲) 的趋势。 超大型液压挖掘机主要用于各种大规模露天矿山的开采及大型基础建设,同时还被用于填海造地工程及港湾河道疏通工程,其中正铲式挖掘机占大部分。目前生产超大型挖掘机的厂商主要分布在美国、日本、德国等工程机械制造强国,著名的生产企业有利勃海尔、卡特彼勒、日立、小松、特雷克斯(O&K)等公司。 1987年,当时世界最大的日立EX3500型正铲超大型挖掘机(斗容量18.8m3,工作重量330t)在日本开发成功,并在这一年又开发了斗容量 22m3,工作重量420t的EX4200型正铲超大型挖掘机。此时世界其他公司也相继开发生产出超大型液压挖掘机,如小松PC1600型、卡特彼勒 5130型,利勃海尔R992型、R994型,特雷克斯(O&K)RH20C型、RH90C型等。 1990年以后,超大型液压挖掘机的斗容量不断增加,达到了50m3以上;机重不断加大,达800t以上,目前已经形成超大型矿山型与土方型液压挖掘机系列。 日立建机于2004年成功研发面向矿山用户的EX8000超大型液压挖掘机,铲斗容量为40m3,大约相当于6辆11吨翻斗车的容量,最大工作重量 780t,是目前日本制造的最大级别的液压挖掘机。该超大型液压挖掘机不仅可顺应为降低采掘运输成本而不断加大矿山自卸车载荷的趋势,而且与钢索式电动挖掘机相比,超大型液压挖掘机具有更好的机动机和操作性,在降低运营成本方面也极具优势。至今,日立建机公司已经向北美、澳大利亚、亚洲和南非等大型矿山交付了大约700台多种型号的超大型液压挖掘机,在该级别中占据了全球最大的市场份额。 超大型液压挖掘机的快速发展与应用主要得益于其具有良好的工作性能和耐久性、可靠性,极大地提高了挖掘机工作效率和经济效益。 优越的操纵控制性及作业性能是超大型液压挖掘机的显著特点。超大型液压挖掘机采用发动机—油泵电气控制系统,可以最大限度地发挥发动机的性能、降低油耗,使机械高效地工作。大部分超大型挖掘机采用故障检测系统,利用装设在驾驶室的显示器为驾驶员提供各主要部件、系统及整机的工作,机械操作的自动化程度及适应性大大提高。 另外,超大型液压挖掘机的工作装置也有了极大的改进,在斗杆与铲斗之间采用了平行伸缩机构,使铲斗的工作范围加大,作业性能提高。为
⑴有利二物料的自由流动,因此铲斗内壁不宜设置横向缘,棱角等斗底的纵向剖面形状要适合各种物料的运动规律。 ⑵要使物料易于卸净,用于粘于的铲斗装卸时不易卸净,因此延长了作业循环时间,降低了有效斗容量,因此可采用强制卸土板的粘土铲斗。 ⑶为了便于物料的铲装,装入物料不易掉出,铲斗宽度与物料颗料直径之比应大于4:1,当此比值大于50:1时颗粒尺寸的影响可不考虑,视物料为均质。 ⑷装设计齿有利于增大铲斗与物料刚接触时的挖掘线比压,以便函切入或破碎阴力较的物料。 2.2 回转机构方案设计 单斗液压挖掘机回转机构的回转时间约占整个工作循环时间的50~40%。回转液压油路的发热量约占液压系统总发热量的30~40%。因此合理地确定回转机构的液压油路和结构方案,正确选择回转机构诸参数,对提高生产率,改善司机的劳动条件,减少工作装置的冲击等具有十分重要的意义 2.2.1 对回转机构的基本要求 1)在加速度和回转力矩不超过允许值的前提下,应尽可能缩短回时间。在回转部分惯性已知的情况下,角加速度的大小变最大的回转扭掩蔽的限制该扭矩不应超过行走部分与地面的附着力矩。 2)回转机构的动载荷系数不应超过允许值,非全回转的挖掘机回转时,工作装置不应碰撞定位器。 3)回转能量损失最小 2.2.2 回转机构的选择 回转机构分为全回转式和半回转式,悬挂式液压挖掘机通常采用半回转的的回转机构,回转角度一般等于或小于180o本次设计为履带式液压挖掘机,因此采用全回转式回转机构。 全回转的回转机构,按液动机的结构形式可分为高速方案和低速方案两类。由高速液压马达经齿轮减速成器带动回转小支承上的回定齿圈动,促使转台转的称为高速方案。其低速大扭矩液压马达直接带支回转小齿轮促使转台回转的称为低速方案。在高速方案中,通常采用行星齿轮减速成箱减速。行星齿轮减速成箱,虽然加工要求较高,但可用一般渐开线齿廓的模数铣刀进行加工,结构也比较紧凑,速成比大,受力好。因此获得广泛的应用,高速方案一般采用斜轴式液压马达驱动的回转机构。此方案结构比较紧凑,速比大,受力好,回此在液压挖掘机中获得广泛的应用,本次设计将采用斜轴式液压 马达驱动的回转机构。高速方案还有以上优点:高速成液压马达体积小,效率高,不需要背压补油,便函于设置小制支器发热和功率损失小,工作可靠。 2.2.3 制动器的选择 制动器应满足工作安全可靠,制动平稳,操纵轻便灵敏等要求,在对回转机构采用了钳盘式制动器。因为通过试验与蹄式制动器相比,它具有下优点: 1)钳盘式制支器所产生的制动力矩比较半稳由于它无增力作用其效率系数K与磨擦系数u为直线关系。 2)由于制动盘都暴露在外,因此通风良好,旋转时还有自洁作用,热稳定性好,基本上无热衰退现象,在连续多次使用情况下制动力矩变化很小,甚至在恶劣工况下仍能正常使用。 3)维修方便,不需要经常调整间隙。 4)制动磨擦衬片磨擦均匀,使用寿命也比较长。但其也存在一些缺点,需采取措施减小及至消除。 2.3 液压系统方案设计 按照挖掘机各个机构和装置的传动要求,把各种液压元件用管路有机地连接想来的组合体叫做挖掘机的液压系统,液压系统的功能是把发动机的机械能以油液为介质,利用液压泵转变为液压能进行传送,然后通过液压缸和液压马达等执行元件转返为机械能实现各种动作。2.3.1 单斗液压挖掘机的作业过程包括下列几个间歇动作