[of] the ancillary system for hydraulic hybrid vehicle

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Theses and Dissertations
2010
Design and control [of ] the ancillary system for
hydraulic hybrid vehicle (HHV)
Mohamed E. Abdelgayed
The University of Toledo
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Abdelgayed, Mohamed E., "Design and control [of] the ancillary system for hydraulic hybrid vehicle (HHV)" (2010). Theses and
Dissertations. Paper 771.
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A Thesis
entitled
Design and Control the Ancillary System
For the Hydraulic Hybrid Vehicle (HHV)
by
Mohamed E Abdelgayed
Submitted to the Graduate Faculty as partial fulfillment of the requirements
for the Master of Science Degree in Mechanical Engineering
________________________________
Dr. Walter W. Olson, Committee Chair
________________________________
Dr. Cyril Masiulaniec, Committee Member
________________________________
Dr. Sorin Cioc, Committee Member
________________________________
Dr. Patricia Komuniecki,
Dean College of Graduate Studies
The University of Toledo
August 2010
Copyright © 2010, Mohamed Ezzat Abdelgayed
This Document is copyrighted material. Under copyright law, no parts of this document
may be reproduced without the expressed permission of the author.
An Abstract of
Design and Control the Ancillary System
For the Hydraulic Hybrid Vehicle (HHV)
by
Mohamed E Abdelgayed
Submitted to the Graduate Faculty as partial fulfillment of the requirements
Master of Science in Mechanical Engineering
The University of Toledo
August 2010
The hybrid hydraulic vehicle (HHV) is a new technology that uses hydraulic
power in conjunction with the conventional vehicle internal combustion engine (ICE) in
order to improve fuel economy for road vehicles propulsion. In addition to propulsion, a
portion of the hydraulic power can be used to drive hydraulic accessories, through a
power take-off point. A power take-off method with HHV decreases the cost of
implementation on the vehicle as the main power source is readily available. The
transferred power, with the appropriate interface using controlled hydraulic circuit, drives
the accessories system.
This research preliminary analysis investigates three proposed systems: hydraulic
intensifier, hydraulic transformer, and pump/motor configuration or hydrostatic system.
The research studies the systems efficiencies and the impact of these systems on the main
hydraulic circuit. The impact is measured by the system ability to isolate the hydraulic
accessories fluid from the main circuit fluid in order to maintain the main circuit
iii
performance. The hydraulic intensifier and the hydraulic transformer do not isolate the
main source fluid from the load fluid effectively. The estimated efficiency is around 55%
for the hydraulic intensifier while it is around 90% for the hydraulic transformer.
Accordingly, both of them are not suitable for the application.
A hydrostatic transmission or the pump/motor configuration provides complete
isolation as the load fluid is separated from the source fluid. In addition efficiency
expected to be in range of 80% to 90%. Consequently, the hydrostatic transmission
system design is used in this application.
A comprehensive analysis is performed on the hydrostatic transmission. The
analysis starts by defining two pressure working ranges from hydraulic tools
manufacturer’s datasheets; the low pressure at 2,000 psi and the high pressure at 10,000
psi. After selection of loading pressure, the design of the driving pumps, and hydraulic
motor with the control valve is performed. The analysis includes controller design to the
system in order to maintain the load demand.
The hydrostatic transmission system design is modeled and then simulated using
MATLAB/SIMULINK. The model simulation incorporates several loading cases in order
to define the time required to drive the hydraulic accessories efficiently. The model result
reports the maximum operating time and the input power consumption rate for each load
case. The input power consumption rates and the assumed efficiencies are verified by
comparing to the manufacturer’s datasheet values and to the model output efficiencies.
Conclusions and recommendations are provided at the end of this research.
iv
To my father who left this world, my lovely mother, my brothers: Mostafa and
Mo’men, and my uncle Yeheya. Without my family members’ love, support, belief and
hope I would not reach this point in my life.
ii
Acknowledgements
First, I would like to thank god for the family I have, people I met, and the
knowledge I gained.
I am very grateful to Dr. Walter Olson for letting me working under his
supervision. Dr. Olson supervision and guidance boost the qualities of my engineering
and research skills. Without his guidance and persistent help, this dissertation would not
have been possible. I am lucky to be one of his students.
Special thanks to the distinguished faculty members who served on my
committee: Dr. Sorin Cioc and Dr. Cyril Masiulaniec. In addition, I would like to thank
Dr. Mohamed Samir Hefzy, Dr. Munir Nazzal, Dr. Eman Mohamed, Dr. Mohamed Abu
Haiba, Dr. Mingwei Shan, Dr. Matthew witte, and Dr. Amr Zaky for their continuous
support, encouragement, and useful suggestions.
I am also grateful for the support and friendship of the members of the hydraulic
hybrid vehicle group Boya, Chao, and Zach.
vii
Table of Contents
Abstract ........................................................................................................................................... iii
Acknowledgements ........................................................................................................................ vii
Table of Contents .......................................................................................................................... viii
List of Tables ................................................................................................................................... x
List of Figures ................................................................................................................................. xi
Nomenclature .................................................................................................................................xiii
Chapter 1: INTRODUCTION.......................................................................................................... 2
1-1 Research background ............................................................................................................. 2
1-2 Problem statement ................................................................................................................. 4
1-3 Work outline .......................................................................................................................... 4
Chapter 2: LITERATURE REVIEW .............................................................................................. 6
2-1 Hydraulic hybrid vehicle ....................................................................................................... 6
2-1-1 Hydraulic accumulator ................................................................................................... 8
2-2 Accessories power system ..................................................................................................... 8
2-2-1 Hydraulic intensifier:.................................................................................................... 10
2-2-2 Hydraulic transformer: ................................................................................................. 15
2-2-3 Pump/Motor configuration: .......................................................................................... 22
2-2-3-1 Pumps .................................................................................................................... 23
2-2-3-2 Hydraulic motors .................................................................................................. 25
2-3 Hydraulic tools .................................................................................................................... 26
Chapter 3 THEORETICAL ANALYSIS ....................................................................................... 28
3-1 Hydraulic Load .................................................................................................................... 31
3-1-1 Pole chainsaw ............................................................................................................... 31
3-1-2 Single acting cylinder ................................................................................................... 33
3-2 Hydrostatic transmission ..................................................................................................... 35
viii
3-2-1 Pump............................................................................................................................. 35
3-2-2 Hydraulic motor ........................................................................................................... 38
3-3 Hydraulic accumulator ........................................................................................................ 40
3-4 System steady state analysis ................................................................................................ 41
3-5 Controller ............................................................................................................................. 46
Chapter 4 SIMULATION MODEL ............................................................................................... 54
Chapter 5 SIMULATION RESULTS ............................................................................................ 61
5-1 Simulation results for case 1 ................................................................................................ 61
5-2 Simulation results for case 2 ................................................................................................ 65
5-3 Summary.............................................................................................................................. 72
Chapter 6 SUMMARY AND FUTURE WORK ........................................................................... 73
6-1 Summary.............................................................................................................................. 73
6-2 Conclusion ........................................................................................................................... 74
6-3 Future work ......................................................................................................................... 75
References ...................................................................................................................................... 76
Appendix A .................................................................................................................................... 79
Appendix B .................................................................................................................................... 81
ix
List of Tables
Table 3- 1 Pumps and loads coefficients ..................................................................... 46
Table 3- 2 Pumps controller coefficients ..................................................................... 53
Table 5- 1 Simulation summary................................................................................... 72
Table A- 1 Maximum operating time for different cases ............................................ 79
Table B- 1 Hydraulic tools input power requirement .................................................. 81
x
List of Figures
Figure 1- 1 Propulsion system configuration a) Conventional b) Series Hydraulic Hybrid
(SHH) configuration .............................................................................................. 3
Figure 2- 1 HHV power take-off overall system configuration ..................................... 6
Figure 2- 2 Series hydraulic hybrid (SHH) vehicle configuration ................................ 7
Figure 2- 3 Maxtor Jet hydraulic intensifier schematic circuit ................................... 11
Figure 2- 4 Modified hydraulic intensifier internal structure ...................................... 12
Figure 2- 5 Hydraulic intensifier circuit ...................................................................... 13
Figure 2- 6 Hydraulic intensifier circuit- case (1) ....................................................... 14
Figure 2- 7 Hydraulic intensifier circuit- case (2) ....................................................... 14
Figure 2- 8 Hydraulic transformer principle Vs. throttling ......................................... 16
Figure 2- 9 Hydraulic transformer external structure ................................................. 17
Figure 2- 10 Hydraulic transformer internal structure ................................................ 18
Figure 2- 11 Hydraulic transformer internal structure section A-A ........................... 18
Figure 2- 12 Hydraulic transformer operation ............................................................ 20
Figure 2- 13 Open loop hydrostatic transmission ........................................................ 22
Figure 2- 14 Gear pump .............................................................................................. 24
Figure 2- 15 Swash plate piston pump ........................................................................ 24
Figure 2- 16 Variable displacement axial piston bent axis pump ............................... 25
Figure 2- 17 Hydrostatic transmission for HHV accessories....................................... 26
Figure 2- 18 Hydraulic tools a) 12 ton Remote Compression C-Head, Huskie Tools b)
General purpose cylinders Rc-Series, Enerpac .................................................... 27
Figure 3- 1 Schematic drawing of hydraulic accessories driving circuit ..................... 29
Figure 3- 2 Pole chainsaw C25 ................................................................................... 31
Figure 3- 3 Single acting cylinder ................................................................................ 33
Figure 3- 4 Eaton series 26 gear pumps ...................................................................... 36
Figure 3- 5 Series 26 gear pump [0.4 in3/rev] ............................................................ 37
Figure 3- 6 Enerpac ZE-3 pumps ................................................................................ 38
Figure 3- 7 Overall system simulation model .............................................................. 47
Figure 3- 8 Circuit operation flow chart ...................................................................... 48
Figure 3-9 Motor controller block diagram ................................................................. 50
xi
Figure 3- 10 Motor circuit block diagram.................................................................... 51
Figure 3- 11 Low pressure pump control circuit ......................................................... 51
Figure 3- 12 ENERPAC pump control circuit ............................................................. 52
Figure 4- 1 Overall system simulation model .............................................................. 54
Figure 4- 2 Low-pressure pump 2,000 psi simulation model ...................................... 55
Figure 4- 3 ENERPAC pump simulation model.......................................................... 55
Figure 4- 4 Quick extend stage pump 1,500 Psi simulation model ............................. 55
Figure 4- 5 High-pressure pump 10,000 pump simulation model ............................... 56
Figure 4- 6 Hydraulic motor simulation model ........................................................... 56
Figure 4- 7 Hydraulic accumulator simulation model ................................................. 57
Figure 4- 8 Load torque simulation model .................................................................. 57
Figure 4- 9 Control valve simulation model ................................................................ 58
Figure 4- 10 Controller simulation model ................................................................... 58
Figure 4- 11 Low-pressure pump 2,000 psi controller simulation model.................... 59
Figure 4- 12 ENERPAC pump 1,500/10,000 psi controller simulation model ........... 59
Figure 4- 13 RPM controller simulation model ........................................................... 60
Figure 5- 1 Output pressure VS. motor speed for case 1 ............................................. 62
Figure 5- 2 Control valve input current and output swash plate angle case 1 ............. 63
Figure 5- 3 Motor output torque and pumps input torque case 1................................. 64
Figure 5- 4 Accumulator outputs case 1 ...................................................................... 65
Figure 5- 5 Accumulator final state case 1 .................................................................. 65
Figure 5- 6 Output pressure VS. motor speed for case 2 ............................................. 66
Figure 5- 7 Control valve input current and output swash plate angle case 2 ............. 67
Figure 5- 8 Motor output torque and pumps input torque case 2................................. 67
Figure 5- 9 Accumulator outputs case 2 ...................................................................... 69
Figure 5- 10 Accumulator final state case 2 ................................................................ 69
Figure A- 1 Loading pressure Vs. maximum operating time for Eaton pump ............ 80
xii
Nomenclature
AL
Load orifice area
AP
Piston area
Ar
Rod area
Cd
Discharge coefficient
Cs
Laminar leakage flow
Cst
Turbulent leakage flow
Cv
Viscous loss coefficient
Cv
Viscous loss coefficient
Dm
Motor volumetric displacement
DP
Pump volumetric displacement
FL
Load applied force
Fo
Spring initial force
FP
Piston force
Fr
Rod force
HHV
Hydraulic Hybrid vehicle
i
Control signal current
k
Spring constant
Kc
Pressure flow coefficient
xiii
kg
Gas specific heat ratio
Kq
Flow gain coefficient
l
Length of the conduit
M
Number of loads
n
Polytropic coefficient
P
Pressure
PHH
Parallel Hydraulic Hybrid
PTO
Power Take Off
Pg
Gas pressure
Pm
Motor suction pressure
Qa
Actual pump/motor flow rate
Qi
Ideal flow rate
QL
Load flow rate
Qleakage
Leakage flow rate
Qo
nominal flow rate
Qs
Supply flow rate
R
Case drain port radius
s
Laplace transform
S
Sommerfeld number
SHH
Series Hydraulic Hybrid
Ta
Actual pump/motor torque
Ti
Ideal torque
Tload
Load torque
xiv
Tm
Motor theoretical toque
V
Fluid volume
Va
Air volume
Ve
Effective volume
Vg
Gas volume
VP
Piston velocity
α
Swash plate angle
αm
Motor angular acceleration
βa
Air bulk modulus
βc
Container bulk modulus
βe
Effective bulk modulus
βl
Fluid bulk modulus
ηf
Piston force efficiency
μ
Viscosity
ρ
Fluid density
σ
Dimensionless number
p
Pump shaft rotational speed in rpm
 t _ motor
Motor torque efficiency
 v _ motor
Motor volumetric efficiency
t _ pump
Pump torque efficiency
v _ pump
Pump volumetric efficiency
xv
Chapter 1: INTRODUCTION
An accessories system on a vehicle is any system that is not included in the
propulsion system. The variety of components used depends on the vehicle type. For
commercial vehicles, accessories include but are not limited to the steering system,
braking system, lighting, and windshield wiper. Trucks and special vehicle accessory
systems may include power tools and power take-off (PTO) devices. The type of the
accessories used depends on the application itself. Examples of these tools are nut
splitter, hydraulic wrench, and hydraulic rams.
1-1 Research background
Hydraulic accessory systems can be powered either manually (hand pumps), or by
a separate system installed on the vehicle (hydraulic crane system), or by power take-off.
Power take-off is a method in which portion of the main power system is used to
drive the auxiliary system, or in this case, the accessory system. Power take off can be
mechanical, electrical, or hydraulic depending on the main power source outputs.
Hydraulic power take-off uses the available hydraulic power source in the form of
pressure and flow rate. The transferred power, with the appropriate interface using
2
controlled hydraulic circuit, drives the accessories system. The main power source in this
research is the hydraulic hybrid vehicle (HHV) power source.
HHV is a new model that uses hydraulic power as the main source for driving the
vehicle; it has two configurations, parallel hydraulic hybrid model (PHH) and series
hydraulic hybrid model (SHH). HHV models have shown a great potential regarding
high fuel economy and flexibility in engine control, especially SHH [1] [33] [13]. Figure
1-1 shows configuration for both conventional vehicle and series hydraulic hybrid vehicle
Figure 1- 1 Propulsion system configuration a) Conventional b) Series Hydraulic Hybrid
(SHH) configuration [1]
In the SHH there is not any mechanical link between the engine and the wheels.
This gives the advantage of operating the engine at the best efficiency point achievable
regardless the speed of the vehicle or the road friction forces. SHH uses hydraulic power
in conjunction with an internal combustion engine (ICE) to drive the vehicle.
3
Introduction of the hydraulic power to the main power increases the systems overall
efficiency [1].
Power take off method with HHV decreases the cost of implementation on the
vehicle as the main power source is readily available. Consequently, this method reduces
the number of components used; hence increasing the overall efficiency. In contrast, the
volumetric efficiency decreases as more hydraulic components are added to the system.
1-2 Problem statement
The main objective of this research is to investigate the available driving systems
that can power the HHV accessories through the power take off. In addition, study the
impact of these systems on the main power source. The impact is measured by the
amount of power drawn, the effect on the performance, and the overall efficiency.
The investigation started with a preliminary study of the available systems then a
comprehensive analysis and design of the most efficient system were conducted. After
studying the selected system, a simulation model was presented using Matlab/Simulink
version R2008b to show the impact on the main power source. The simulation model
included the main power system, the driving system, and the hydraulic tools.
1-3 Work outline
Chapter two presents the literature review. This includes a preliminary study for
the suggested systems.
Chapter three presents the circuit analysis, considerations that were taken when
designing the circuit, and design assumptions.
Chapter four presents the simulation model using Matlab/Simulink software.
4
Chapter five presents and discusses the simulation results including the power
used, the operation cycle time, and the impact on the main system.
Chapter six is the conclusion of the model results. In addition, it presents the
future research work in order to improve the accessory system.
5
Chapter 2: LITERATURE REVIEW
Accessories power system is a hydraulic circuit that uses HHV power system
through power take-off to operate hydraulic accessories, or hydraulic tools. The process
should be performed in a highly efficient manner with the least impact on HHV power
system. Figure 2-1 shows the overall system configuration.
Figure 2- 1 HHV power take-off overall system configuration
2-1 Hydraulic hybrid vehicle
HHV power is the main power source to drive the hydraulic tools. To illustrate
the operation of HHV, figure 2-2 shows the configuration of the SHH.
6
Figure 2- 2 Series hydraulic hybrid (SHH) vehicle configuration [1]
SHH model consists of a high pressure accumulator, low pressure accumulator,
pump/motor on each axle, and a pump/motor that is connected to the internal combustion
engine (ICE). The operation of the SHH starts with, when there is sufficient pressure in
the accumulator, the pressurized fluid is directed to operate both pumps mounted on the
axles; the engine is in off mode. Once the pressure falls to a predefined limit, the fluid is
directed to start the engine through the pump/motor unit (P/M) as it works as a motor at
this stage. Once the engine starts, the pump/motor returns to pump mode to drive the
axles and charge the accumulator. When the accumulator is charged, the cycle is repeated
and the accumulator becomes the driving source. SHH is designed to capture energy,
when braking, which is limited by capacity of the accumulator; this energy is referred to
as regenerative braking energy. Several researches were devoted to implement the HHV
models; with engine off mode, regenerative braking energy, and sophisticated control
techniques SHH showed more reduction in fuel consumption compared to the
conventional vehicle, up to 70% [15][13][33][1].
7
2-1-1 Hydraulic accumulator
Hydraulic accumulator is categorized as a hydraulic accessory; it is used to store
hydraulic energy. It is characterized by high energy density and accepting high rates of
charging and discharging [1]. There are different types of accumulators; examples are gas
charged, spring loaded, and gas over bladder. All accumulators’ pressure decreases as
fluid discharges except weight-loaded accumulator. Weight loaded accumulator
maintains pressure until all oil is used [35]. Hydraulic accumulators are used in many
applications such as supplementing pump flow in circuits with medium to long delays
between cycles, or hold pressure in a cylinder while the pump is unloading or stopped.
Also, they can be used as a ready supply of pressurized fluid in case of power failure, or
reduce shock in high velocity flow lines or at the outlet of pulsating piston pumps. High
pressure hydraulic accumulator is the source for the hydraulic power in HHV. The power
used by the hydraulic accessories system cannot exceed the available power by the high
pressure accumulator. Also, the time for using this power should be within the time taken
to discharge the accumulator at a given pressure and flow rate [35] [10]. In this research,
hydraulic accessories system is used to discharge the high-pressure accumulator; the
HHV main circuit is responsible for charging the accumulator.
2-2 Accessories power system
Accessories power system is the system that uses the HHV vehicle power,
through the power take off point and with appropriate control system, to drive the
hydraulic tools. There are two types of controlled hydraulic system; valve controlled
hydraulic systems and pump controlled hydraulic systems [11] [2].
8
Valve controlled hydraulic systems use valves to transmit power from the power
source, or the power take off, to the load which is the hydraulic tools. The type of valve
used depends on the application, system requirements such as stability, efficiency, and
type of load. Load, or actuator, can be either translational as a piston cylinder or
rotational as a hydraulic motor. Valve controlled systems are used when the requirements
are either quick response, less bulky systems, or when using multiple loads from one
source at a time [11] [2]. The disadvantages of these systems are low efficiency, and their
higher cost. The high cost stems from the additional components required for heat
exchanging to dissipate energy wasted; the heat is generated when the pressure drop
across the valve becomes high. Example of this system is hydraulic intensifier.
Pump controlled systems use the pump to drive the load, or the hydraulic tools;
the pump power source is a mechanical power source to drive the pump shaft. Pumps can
be either fixed displacement or variable displacement pump. For the fixed displacement
pump, the output power is controlled by controlling the shaft speed only. In the variable
displacement pump, the output power is controlled either by controlling the shaft speed,
the pump volumetric displacement, or both, which gives more capabilities to operate
different load requirements. Pump controlled systems are characterized with high
efficiency and higher power outputs ranges. The disadvantages of these systems are
slower response than valve controlled systems, more space to operate for most
applications, and a single pump cannot be used to operate different loads from the same
power source [11][2]. Example of this system is hydrostatic transmission.
Hydraulic Transformer is another system that has the characteristics of both
systems, pump controlled systems and valve controlled systems. It has the combination of
9
the pump and the valve [28] [34]. A Hydraulic Transformer has the advantages of the two
systems and some of their disadvantages. It is characterized with high efficiency and
quick response but it cannot operate different loads from the same input power source.
2-2-1 Hydraulic intensifier:
Hydraulic Intensifier is a hydraulic valve used to amplify pressure with a ratio
equal to the reduction in area according to Pascal’s law (Eq. 2-1, 2-2). The internal bore
has differential diameter where the large end is connected to the source pressure and the
output is connected to the output circuit or in this case, the hydraulic tool [22].
F  P* A
(2-1)
Pr AP

PP Ar
(2-2)
Maxtor Jet Company [21] uses this principle to increase the pressure of a water
source. The intensifier then pumps the pressurized water into the user’s facilities. Figure
2-3 shows schematic drawing of their circuit.
10
Figure 2- 3 Maxtor Jet hydraulic intensifier schematic circuit [21]
Maxtor Jet’s intensifier uses the hydraulic circuit input pressure and amplifies it
with ratio equals to the piston differential area. The output pressurized fluid then is
pumped into the accumulator that feed the load. The accumulator is used to smoothen the
load input flow and isolate it from pulsating flow that is caused by the intensifier
reciprocating action.
As mentioned the main points to evaluate the intensifier adequacy to the HHV
circuit are the impact on the main circuit and intensifier efficiency.
When the fluid used to drive the hydraulic tools is the same as the hydraulic units,
the possibility of contaminating the fluid is high; in turn, this affects the performance of
the accessories circuits and accordingly affects the main power circuit efficiency. The
contamination may include solid particles or entrained air. Entrained air has the effect of
reducing the fluid capability to transmit power. It can also wear the hydraulic component
11
internal structure or cavitation, which has damaging effects on components [11] [36].
Based on this analysis, modifications should be made to ensure isolation.
There are two possible ways to ensure isolation, either by modify the intensifier
structure, or by modifying the hydraulic circuit. The problems with the latter option are,
the modification will include adding more components for isolation and this will decrease
the volumetric efficiency. The second reason is the economic perspective; with the
increased number of components the system will be more expensive; so modification to
the intensifier structure is more favorable.
To isolate the main source from the load, the load (auxiliary) fluid source should
be different, or isolated, from the main circuit fluid source. Moreover, the intensifier has
to separate both fluids when it is working; so internal and external isolations are required.
For internal isolation, a modification is made to the intensifier internal structure;
Figure 2-4 shows the modified intensifier internal structure.
Figure 2- 4 Modified hydraulic intensifier internal structure
In the modified intensifier, there are two leakage ports: one for the vehicle fluid
source which is close to the power side and the other for the output fluid that is close to
the load side. Each fluid leakage source is captured in a separate compartment and then
12
rerouted to a different tank; these compartments are separated by the middle
compartment, in case of contamination, not to return back to either source.
To show external isolation figures 2-5, 2-6, and 2-7 show the hydraulic intensifier
circuit at idle and working cases in both sides respectively. Each of the HHV fluid
sources is different from the accessories load to ensure isolation.
Figure 2- 5 Hydraulic intensifier circuit
13
Figure 2- 6 Hydraulic intensifier circuit- case (1)
Figure 2- 7 Hydraulic intensifier circuit- case (2)
To illustrate the hydraulic intensifier operation, the intensifier working case (1),
figure 2-6, is discussed in detail. Starting from HHV power source, the fluid discharges to
intensifier inlet port through the directional control valve (DCV). The pressurized fluid
acts on the larger piston area causing the piston to move to the left. Meanwhile, this
action results in suction of the accessories oil from the accessories source tank due to the
14
movement of the smaller piston from the right side through the check valve (D). On the
left side, the other small piston tends to pressurize the accessories fluid and pumps it to
the accumulator using the check valve (A). The output from the accumulator supplies the
hydraulic tools or accessories and the low pressure fluid from the accessory returns back
to the accessories tank. In this case, the check valves (B and C) are closed. The same
cycle is repeated, as shown in case (2), in the reverse direction using the DCV to change
the flow direction. In case (2) the check valves (A and D) are closed.
After studying adequacy and modifying the intensifier structure to reduce the
impact on the main circuit, the next criterion, to evaluate the intensifier, is efficiency. To
estimate the efficiency, a comprehensive study has to be conducted on the intensifier
which is not necessary in this preliminary analysis. With the aid of reports from several
intensifier manufacturers [23] [24] [25], the efficiency range reported is from 25% to
95%. These values depend on the working condition, the higher pressure working
condition the higher the intensifier efficiency. These reports didn’t include the DCV
effect on efficiency. If a 4 way open centered DCV with rectangular port geometry with
load pressure designed as PL=Ps/2 is used, the efficiency is 0.125, with the load port
opening one quarter of the maximum opening. (PL is load pressure and Ps is supply
pressure ) [11].
2-2-2 Hydraulic transformer:
Hydraulic Transformer is a hydraulic device that transforms fluid hydraulic
power, in form of pressure and flow rate, into a different pressure and flow rate with
minimum power loss. Figure 2-8 shows the principle of the hydraulic transformer.
15
Figure 2- 8 Hydraulic transformer principle Vs. throttling [29]
Power transformation minimizes the power losses; as the pressure is transformed
to a lower level with increase in the flow rate, a third flow connection QT is added, to
balance the flow equation. In contrast, throttling the fluid results to energy loss as the
pressure decreased for the same flow rate; the pressure decreases due to friction.
Starting the transformer operation with the transformer connections, figure 2-9
shows the transformer external structure. The transformer has three ports; one is
connected to the high-pressure line PH, the second port is connected to the load PN, and
the third is connected to the low-pressure line PL.
16
Figure 2- 9 Hydraulic transformer external structure [34]
Hydraulic transformer consists of a rotating shaft that is connected to the piston
block through splines. The piston block contains nine pistons that are placed
symmetrically around the piston block center. Each piston is connected to the swash plate
by which its angle determines the piston stroke. The input ports: high pressure, low
pressure, and load port are connected to the piston block through the stationary sealing
plate, face plate, and rotating sealing plate. The face plate is mounted on the shaft through
bearing. The face plate angular position determines the load output pressure. The
adjusting mechanism, that is connected to the face plate with pinion gears, changes the
face plate angular position. The rotating plate is connected to the shaft through splines; as
the shaft rotates, the rotating sealing plate rotates. The transformer internal structure and
section A-A are shown in figures 2-10 and 2-11 respectively.
17
Figure 2- 10 Hydraulic transformer internal structure [34]
Figure 2- 11 Hydraulic transformer internal structure section A-A [34]
The hydraulic transformer is effectively both a motor and a pump. As with
conventional swash plate hydraulic motors, high-pressure hydraulic fluid is used to
18
produce shaft rotation. In turn, this shaft rotation is used to pump hydraulic fluid (at a
lower pressure than that used to motor) into hydraulic load circuit. Thus, the hydraulic
transformer is very efficient pressure divider. This is accomplished by multi-partitioned
valve plate. On a conventional swash plate hydraulic pump or motor, the valve plate
consists of two kidney shaped orifices: one for suction and the other for discharge. In a
hydraulic transformer, the valve is sectioned into three or more orifices: one for suction,
one for discharge, and the remaining for orifices connected to the hydraulic load(s).
Figure 2-12 (directly from [34] figure 4-9) shows a three section valve plate with
respect to piston locations and swash plate position. The swash plate fixed to the
transformer body forces a sinusoidal height on the pistons over one complete rotation of
the barrel assembly. The valve plate is adjustable with respect to the swash plate but once
adjusted is fixed in location with respect to the body. As in a hydraulic motor, rotation of
the barrel assembly is caused by applying high-pressure fluid (PH) into the piston cylinder
when it approaches the top dead center. This causes a force vector tangent to the barrel
cross section at the pitch circle of the pistons producing a torque on the barrel.
19
Figure 2- 12 Hydraulic transformer operation [34]
As shown in Figure 2-12 a, the barrel with the pistons rotate under high pressure
(PH) opening occurs near 0º shaft angle and closes near 120º (the total opening is
somewhat less to allow full valve closure). It is assumed that valve opening are of equal
length and that the valve ribs or unopened parts are sufficient to achieve full valve shut
off without cross porting. After the high-pressure fluid is shut off, the further rotation of
the barrel causes the pistons to slide under the low-pressure (PL) opening of the valve
plate and discharges oil until the rotation of the shaft at angle near 240º. Because there is
20
remaining volume of fluid in the piston cavity, further rotation passes the remaining fluid
under the load opening of the valve plate and the upward force caused by the piston
exhausts oil under an externally controlled demand pressure (PN). Near a shaft rotation
angle of 360º, the barrel rotation rotates the piston opening away from the load to the
high-pressure port repeating the cycle.
Figure 2-12 a shows the valve plate angle at 60º where high pressure oil is entered
the piston cavity at the top dead center accompanying a shaft rotation angle of 0º. As a
result, the demand oil pressure (PN) and flow can be maintained nearly equal to the high
pressure PH and transformer suction flow with minimal losses. Figure 2-12 b shows the
valve angle shifted 30º thereby causing high-pressure fluid to enter the piston cavity
before top dead center. This causes the load pressure PN to be less than the high pressure
PH. However, the flow is greater as makeup oil is brought in from the transformer
discharge port. In the most extreme case shown in figure 2-12 c, the valve plate is rotated
to position 0º (60º away from the position at figure 2-12 a). Accordingly, the highpressure port has little effect on the transformer. The demand oil pressure P N and flow is
nearly the same as the low pressure PL on the transformer discharge side.
To evaluate the hydraulic transformer, the impact on the main circuit and the
efficiency will be discussed. Hydraulic transformer does not provide isolation between
the main circuit and load as the source and the load ports are directly connected in one
device. However, this problem can be solved by using a separate circuit of a reservoir and
pump/motor configuration to drive the transformer but this is not the main goal of the
research plus adding more components reduces the efficiency.
21
Hydraulic transformer has high efficiency up to 90% as reported by Innovative
associates INNAS [28]. However, it is not practical for driving multiple loads from the
same source as it will require an input pressure and flow-rate variations at each load. As a
result, driving rotary loads requires fast and expensive actuators for changing swash-plate
angle.
2-2-3 Pump/Motor configuration:
The last proposed system is pump and motor configuration or hydrostatic
transmission. Figure 2-18 shows a schematic drawing of open loop hydrostatic
transmission circuit where pump, when the shaft rotates, drives the hydraulic motor. The
directional control valve (DCV) is used to control the motor direction. The solenoid is
used to shift the DCV from its default position which is set by the spring. The relief valve
is used to limit the pressure in the circuit to the pressure set for the pump [8] [17].
Figure 2- 13 Open loop hydrostatic transmission
22
Hydrostatic transmission circuits can be high efficiency circuits if proper
components are selected for each task.
2-2-3-1 Pumps
There are two types of pumps according to energy addition classification,
dynamic and positive displacement [36]. Dynamic pumps are characterized that energy is
continuously added to the liquid to increase its velocity, hence the flow rate; the bestknown example are centrifugal pumps. Positive displacement pumps, such as piston
pumps, are the ones that energy is added periodically by the direct force acting on the
moving element and it is independent of the applied pressure.
Positive displacement pumps are recommended over dynamic pumps when the
application requires high-pressure load, high efficiency, or constant flow with variable
pressure system [11] [2] [36]. For this research as high efficiency, high pressure, and
variable pressure is required, positive displacement pumps will be used. Examples of
positive displacement pumps are gear pumps, bent axis piston pump, and axial piston
swash plate pumps.
Gear pump, figure 2-14, has two gears meshed together and as they rotate the inlet
fluid is pressurized due to squeeze action. Gear pumps efficiencies are in the range of
80% to 90% [11].
23
Figure 2- 14 Gear pump [30]
Axial piston swash plate pumps, figure 2-15, consist of several pistons in a
cylinder block; the output flow rate is controlled by the swash plate angle, which is
controlled by another mechanism, usually a servo-valve. Efficiencies are in the ranges of
85% to 95% [11] [35].
CYLINDER BLOCK
SWASHPLATE
SHOE PLATE
VALVE PLATE
ROTATING SHAFT
INLET PORT
OUTLET PORT
PISTON
RETRACTER SPRING
PISTON SHOE
Figure 2- 15 Swash plate piston pump [31]
Bent axis piston pumps, figure 2-16, consist of a cylinder block that is at an angle
to the drive shaft, it is connected to the drive shaft with universal joint. The cylinder
24
block angle is controlled by variable position mechanism. Efficiencies are in the ranges
of 85% to 98%; however the pump design is the most complex pump design and the unit
is expensive comparing to other pumps.
CYLINDER BLOCK
VALVE PLATE
THRUST PLATE
ROTATING SHAFT
UNIVERSAL JOINT
VARIABLE POSITION
MECHANISM
PISTON
Figure 2- 16 Variable displacement axial piston bent axis pump [31]
2-2-3-2 Hydraulic motors
Usually most of the positive displacement pumps can work as a hydraulic motor
with little to no modifications. It is preferable to use swash plate piston type as they have
low starting torque [11] [2] [17].
Based on the previous comparison, piston type pumps and motors show high
efficiency comparing to the other type pumps; consequently, these will be included into
the proposed circuit.
25
Before evaluation the hydrostatic transmission circuit impact on the main system,
the proposed circuit should be introduced. Figure 2-18 shows the proposed circuit to
drive the hydraulic tools.
Figure 2- 17 Hydrostatic transmission for HHV accessories
In the proposed circuit, the high-pressure fluid from the HHV side is used to drive
the variable displacement swash piston motor. The motor shaft drives both the highpressure and low-pressure pumps. Each pump is used to drive its counterpart load. The
fluid returns back to a separate reservoir.
The proposed circuit isolates the hydraulic accessories fluid source from the main
hydraulic circuit. Also with proper selection to the hydraulic motor and pumps, efficiency
is expected to be high comparing to other systems.
2-3 Hydraulic tools
Hydraulic tools represent the load in this research; the type of the tools considered
defines the working condition. Appendix B shows a comprehensive database for the
available tools that can be installed on a vehicle. From the database, the range of pressure
26
required is from 1500 psi to 10,000 psi. Figure 2-7 shows samples of these tools to be
used .
(a)
(b)
Figure 2- 18 Hydraulic tools a) 12 ton Remote Compression C-Head, Huskie Tools b)
General purpose cylinders Rc-Series, Enerpac [19] [20]
27
Chapter 3 THEORETICAL ANALYSIS
The literature reviewed three different systems that can be used to drive the
accessories system on HHV equipped with power take-off. These systems are hydraulic
intensifier, hydraulic transformer, and pump/motor configuration.
The hydraulic intensifier did not isolate the main source fluid from the load fluid;
however, the problem can be solved with a different internal structure. The modified
internal structure provides isolation, but the structure proposed needs frequent
maintenance to prevent leakage. Intensifier efficiency is up to 95% at working pressure
around 150,000 psi [24]. From the manufacturer’s datasheets, the estimated efficiency is
near 55% for the anticipated working conditions; therefore, it is not suitable for this
application.
Hydraulic transformer’s efficiency is up to 90%, however it requires a separate
circuit that provides isolation as the source and the load ports are directly connected in
one device. Accordingly, it cannot be used in this application.
Hydrostatic transmission or the pump/motor configuration provides complete
isolation as the load fluid is separated from the source fluid. In addition, with proper
design for each component, efficiency expected to be in range of 80% to 90%.
Consequently, the hydrostatic transmission system design is the most adequate system
that can be used to drive the hydraulic accessories.
28
The next step was to design, simulate and control the system that operates the
hydraulic tool from the power take off source using the hydrostatic transmission. The
study reported the amount of power required from the main power source, the maximum
operation cycle time for several cases using the available power source, and the overall
efficiency for the proposed system.
In this chapter, detailed analysis for the proposed circuit will be presented. Figure
3-1 shows a schematic drawing for the circuit.
Figure 3- 1 Schematic drawing of hydraulic accessories driving circuit
As shown in the figure, the circuit connects the vehicle side-the power sourcewith the tool side-the output load. The high-pressure accumulator on the vehicle side
provides high-pressure fluid to the accessories circuit. The flow of the fluid is controlled
by the control valve no.1. In case valve no. 1 is opened, the pressurized fluid drives the
variable displacement hydraulic motor, M. The motor maximum output torque depends
on the available pressure of the high-pressure accumulator fluid. The low-pressure output
is directed back to the vehicle tank or the low-pressure accumulator. The hydraulic motor
shaft is connected to both fixed displacement pumps P1 and P2 through coupling and
29
reduction gear. The reduction gear is used to operate the high-pressure pump P2 at a
speed lower than the hydraulic motor and low-pressure pump P1 speed. Pump P1 has the
capabilities to supply pressure to the hydraulic tools in range up to 2,000 psi, and pump
P2 is used in ranges up to 10,000 psi. The circuit is supported with filter, safety relief
valves, and unloading valves no. 2 and 3. The unloading valves are used to reroute the
fluid to the reservoir in case of unloading in order to minimize the power consumption.
The controller, with three inputs and four outputs, operates the system to meet the load
requirement. The inputs are motor shaft speed from tachometer (speed), low-pressure
pump discharge pressure, and high-pressure pump discharge pressure. The outputs are the
control signals to the control valve, the unloading valves, and the hydraulic motor swash
plate angle positioning solenoid. Pumps circuit for the PTO system has a separate tank to
supply both pumps with fluid. The size of the tank must be large enough to provide the
flow demand and avoid fluid shortage on the suction side.
The analysis started by applying two different hydraulic loads, one with pressure
in range of 10,000 psi and the other with pressure in range of 2,000 psi. From the load
analysis, the pumps minimum requirements and the operation cycle time were
determined. After studying the load, both pumps variables were specified based on load
requirements. The next step was to design the hydraulic motor parameters to operate both
pumps efficiently within the hydraulic accumulator capabilities. After designing all
components, the controller design was conducted.
30
3-1 Hydraulic Load
Hydraulic tools analysis includes two components: pole chain saw as a tool that
requires 2,000 psi and general purpose single acting cylinder as a tool that requires
10,000 psi.
3-1-1 Pole chainsaw
Pole chain saw is a hydraulic tool used in tree pruning and brush cleaning. Figure
3-2 shows model C25 pole chain saw, STANELY hydraulics [16].
Figure 3- 2 Pole chainsaw C25 [16]
As shown, it consists of three parts quick coupler, extension bar, and chainsaw
blade. Quick coupler is used to connect the chainsaw with hydraulic power source.
Extension bar is used to increase the range of using the chainsaw. The chainsaw blade
performs the chainsaw main function.
31
Based on the manufacturer data sheet, the chainsaw requires fluid with pressure
2,000 psi and flow rate from 4 to 6 gpm. The designed flow rate is 4 gpm.
To model the chainsaw a fixed area orifice equation is used [11] [2]
QL  AL * C d
2

(3-1)
*P
The linearized equation for the orifice flow equation using Taylor series
Q
1
Qo  K q x  K c P
2
Kq 
Q A
A x
Kc 
Q
P
o
o

 Cd
2

(3-2)
p
(3-3)
ACd
2 P
(3-4)
For fixed orifice area Kq =0, so the equation becomes
1
Q  Qo  K c P
2
(3-5)
Pressure rise equation in varying closed volume is used to determine the
relationship between the pressure and the flow rate in the chainsaw [11] [2]
P  e

(Qs  QL )
t
V
(3-6)
32
3-1-2 Single acting cylinder
Single acting cylinder is a linear actuator that uses hydraulic power to move a
load. The capacity of the cylinder depends on the material, geometry of the cylinder.
Figure 3-3 shows the single acting cylinder.
Figure 3- 3 Single acting cylinder
As shown, the cylinder consists of inlet port, piston, spring, and rod. The
pressurized fluid enters the cylinder from the inlet port; as the pressurized fluid is acting
on the piston area the cylinder extends. The cylinder rod is connected to the load to
transmit the mechanical force that is generated by the pressurized fluid acting on the
piston. The spring is used to retract the piston when there is no pressurized fluid acting
upon it.
33
The single acting cylinder used is ENERPAC RC-1006 [20], Based on the
manufacturer data sheet, the cylinder requires fluid with pressure no more than10, 000 psi
with maximum force 100 ton and effective piston area 20.63 in2. The flow rate
requirement is based on the application and piston speed requirements. The flow rate is
determined by the equation
QL  AP *V p
(3-7)
The operation of the single acting cylinder consists of two stages; in the first
stage, the cylinder extends rapidly as there is no load acting upon the cylinder other than
the piston-rod weight and the spring force. It is characterized by high flow rate and low
pressure. In the second stage, the piston starts to act upon the load using the pressurized
fluid acting on the piston. This stage is characterized by low flow rates and high pressure.
The load requirement in this research in term of piston speed and pressure is as follows:
for first stage, the designed piston speed is 9 in/min and maximum pressure 1500 psi; for
the second stage, the designed piston speed is 1 in/min at maximum pressure 8,500 psi.
From eq. 3-7, the required flow rate is 0.8 gpm for the first stage and 0.089 gpm for the
second stage.
The required pressure, in stage one, is determined by Newton second’s law
equation
 f PA  mg  kx  Fo
(3-8)
In stage two, the equation becomes
34
 f PA  mg  F  kx  Fo
(3-9)
The spring forces and piston-rod weight are neglected comparing to the applied
force and pressure in the second stage as it is used only for piston retraction.
The inlet port acts as a fixed area orifice with the same equation (3-3, 3-4, and 35) with different coefficients as the pressure and flow rate are different. Also, the pressure
rise equation is the same as (3-6).
3-2 Hydrostatic transmission
Hydrostatic transmission circuit consists of two pumps and a variable
displacement hydraulic motor. The variable displacement motor uses the hydraulic power
through the power take off point and drives the two pumps; these pumps drive their
counterpart load.
3-2-1 Pump
There are two different fixed displacement pumps used to drive the load, one with
pressure up to 2,000 psi and the other with pressure up to 10,000 psi.
For the fixed displacement pump, as the speed changes the theoretical output flow
rate changes according to the equation [11] [37]
Qi   P * DP
(3-10)
35
Both pumps have different volumetric and overall efficiencies depending on the
type of the pump, operating point in terms of speed and pressure. The pump leakage is
determined by Poiseuille’s equation [11] [2] [8]:
Qleakage=
R 4 p
R 4 1
or Qleakage= kp , k=
8 l
8 l
(3-11)
For the 2,000 psi pump, Eaton gear pump was incorporated into the model to
evaluate the power required by the pump [27]. Figures 3-4 and 3-5 show the different
model for series 26 gear pumps and the power input power required for the gear pump
with volumetric displacement 0.401 in3/rev.
Figure 3- 4 Eaton series 26 gear pumps [27]
36
Figure 3- 5 Series 26 gear pump [0.4 in3/rev] [27]
The gear pump with volumetric displacement 0.401 in3/rev was selected for many
reasons. It can pump more than the required flow rate which is 4 gpm at 3000 RPM with
a low required input power. In addition, the pump efficiency is higher than other pumps
efficiency with higher volumetric displacement operates at a lower speed.
For the pump with pressure range of 10,000 psi, the same procedure was
followed. ENERPAC, the manufacturer of the single acting cylinder, recommends the
same brand pump to operate the cylinder. Figure 3-7 shows the characteristic of
ENERPAC ZE-3 pump.
37
Figure 3- 6 Enerpac ZE-3 pumps [38]
The unit has two integrated pumps that operates at 1750 rpm and input power 1
hp; one is a fixed displacement gear pump with volumetric displacement 0.142 in3/rev
and the other is a fixed displacement piston pump with volumetric displacement 0.024
in3/rev. The gear pump operates at the first stage with output flow 250 in3/rev and
maximum pressure 1500 psi with total efficiency up to 80%. The piston pump operates at
the second stage with output flow rate 42 in3/rev at 1750 rpm and maximum pressure
8500 psi with total efficiency up to 90%. A lift check valve piston type [38] separates the
first and second stage pumps outputs. At high-pressure load, the piston check valve
closes the first stage pump discharge port and open the second stage pump discharge port
[3].
3-2-2 Hydraulic motor
The hydraulic motor is hydraulic actuator that uses HHV power source through
the power take off to drive both pumps and consequently the load. It has to produce
38
enough power to drive the circuit. The circuit requirements depends on the pumps
requirements which depend on their counterpart load. Axial piston swash plate motor is
characterized with high total efficiency up to 90%.
The governing equations for the hydraulic motor [11] [2] [12]
I m  Tm  Tload
(3-12)
Tm 

* Dm * Pm
 max
(3-13)
Tload   M DP * Pp
(3-14)
Swash plate angle controls the volumetric displacement and consequently the
flow rate. It depends on the geometry and the motor size. In this research, the maximum
value is 21º. To consider the swash plate angle effect, the swash plate angle was
expressed in term of the maximum swash plate angle as it was scaled from 0 to 1.
Electro-hydraulic Servo valve is used to control the swash plate angle based on
the controller output control signal. This control signal energizes the valve solenoid that
control the valve spool displacement and consequently the swash plate angle. To include
the valve into the model, the valve gain and the time constant have to be evaluated. Based
on Bosh Rexroth models [18] for electro-servo valves, the settling time was selected to be
12ms. As a first order system, the time constant became 3ms. To select the gain, Gain =
Max output/Max input. The maximum output, which is the swash plate angle ratio, was
39
1. For a maximum input current of 3A in order to reach this output, the gain became 0.25.
Based on these values, the valve transfer function was

i

0.25
0.003S  1
(3-15)
To incorporate the hydraulic motor into the model, Eaton axial piston swash
motor was incorporated with the following specifications: variable displacement swash
plate piston motor, max power input 25 hp, volumetric displacement 0.561 in3/rev,
moment of inertia 1.2 lb.in2, and Maximum pressure 5000 psi.
3-3 Hydraulic accumulator
The hydraulic accumulator represents the hydraulic power source; it is mounted
on the HHV side. Charging the accumulator is performed from the HHV circuit not the
PTO circuit; PTO operation depends only on the state of charge of the accumulator. The
accumulator selected was a hydro-pneumatic accumulator model, which consists of a precharged inert gas chamber, and a fluid chamber connected to a hydraulic system. A
bladder with elastomeric foams in the gas-side to reduce heat loss separates the chambers.
The dynamic accumulator model is based on the following equation:
kg
PgVg  C
(3-16)
In this model, the accumulator capacity used is 100 liters. The maximum fluid
capacity is 68.3 liters. The maximum pressure is 5000 psi and the minimum pressure, no
40
fluid case, is 1000 psi. The discharge flow rate is determined by the motor requirements
[13].
3-4 System steady state analysis
Steady state analysis evaluates the system ability to perform what is designed for,
i.e. hydraulic motor is able to drive both pumps and they are capable of operating their
load effectively. This analysis is important before studying system dynamics and
controller design.
Before steady state analysis, some considerations should be addressed; these
considerations include bulk modulus, systems efficiencies, and the hydraulic accessories
system reservoir.
Bulk modulus (β) is a fluid property that indicates fluid elasticity or
compressibility; it is the measure of the change of the pressure with respect to the change
of the volume comparing to the original value (Eq. 3-20). Bulk modulus affects system
performance, dynamics, and the efficiency considerably. [11] [2]
 
P
P
Or   V
V
V
V
(3-17)
The negative sign reflects that the fluid is compressed and the final volume is less
than the initial one.
Entrained air existence in the fluid reduces the effective fluid bulk modulus. This
effect can be illustrated from the following equation [11] [14]:
41
1
e

1
l

Va 1
1

Ve a c
(3-18)
To show the effect of the entrained air more clearly, a numerical example is
introduced. If the oil used is SAE-30 with bulk modulus 1500 MPa [26]; this oil is
transported into hoses with bulk modulus 2005 MPa [11] [9], and the volume of the
entrained air of bulk modulus 0.142 MPa [9] is 0.02 to the effective volume then by
substituting in the previous equation, the effective bulk modulus equals 7.04 MPa.
Therefore, the existence of 2% of air in the fluid reduces the bulk modulus to 0.46% of
the fluid bulk modulus original value.
The effect of the bulk modulus on the system performance is illustrated in
pressure rise equation. [11] [2] [6]
P 
V
 (Q 
)
t V
t
(3-19)
If there is no leakage, at constant flow rate, the pressure rise is function of the
bulk modulus; the greater the bulk modulus, the faster the pressure rise in the system
according to the equations

 P   V Qt
P

V
(3-20)
Q *t
(3-21)
From the previous equations, the bulk modulus value affects the time needed to
reach the applied pressure. At high bulk modulus values, the pressure reaches its final
value nearly instantaneously.
42
The pressure rise can be assumed instantly if the fluid volume V is very low or the
load response time is much higher than the driver time constant. The load response time,
for example the single acting cylinder, is in range of seconds while the pumps response
time is in ranges of milliseconds. The fluid volume is considerably small if the hoses are
short or the driving system is close to the load.
The effect of the bulk modulus on components efficiency is demonstrated from
the following equation [7]:
 v _ pump

Qa
Cs P Cst
 1


Qi
xS 
x
 v _ motor

Qi
Qa
S

P
,

(3-22)
1
P
C
C
1 s 
 st
xS 
x

D1 3
12
 P 
 2

  
,
x
(3-23)
sin 
sin  max
(3-24)
As shown, as the bulk modulus decreased, the pressure term increased causing the
volumetric efficiency to decrease considerably; this is true for both pump and motor.
Driving system and hydraulic tools efficiencies determine the system
performance. The efficiency depends on the operating point; it is not constant for all
points. Pumps and motors volumetric and torque efficiencies can be determined from the
equations [7]:
 v _ pump 
Qa
Cs P Cst
 1


Qi
xS  x
(3-25)
43
 t _ pump 
 v _ motor

t _ motor
S

P
,
Ti
Ta
Qi
Qa


1

1
Cf
Cv S

 C h x 2 2
x
x
1
P
C
C
1 s 
 st
xS 
x
(3-27)
Cv S C f
Ti
 1

 Ch x 2 2
Ta
x
x

(3-26)
D1 3
12
 P 
 2

  
,
x
sin 
sin  max
(3-28)
(3-29)
To obtain efficiencies for any operating point, the previous equations coefficients
should be determined for each unit. To determine these coefficients either a complete unit
data in term of pressure and speed should be available or testing each unit experimentally
over the operating range should be conducted. To put the efficiency into the model, the
system will be tested on the maximum operating point to obtain the maximum power
drawn from the main system and the efficiency included will be the value at this
operating point only, i.e. efficiency is constant. This assumption is valid with the positive
displacement pumps or motors as the efficiency curve is almost flat around the maximum
efficiency and the difference between the maximum and minimum efficiency is around
5%.
The reservoir size is critical in circuit performance as it supplies oil to two pumps,
high pressure and low-pressure pump simultaneously. The fluid should be available
whenever the pumps are in operation. The common practice for the reservoir size is [37]
44
Reservoir size (gal) = 3 X pump flow rate (gpm)
(3-30)
As there are two pumps, the first has a maximum flow rate of 5.2 gpm and the
second has a maximum flow rate of 1.21 gpm then the reservoir size has to be, at least, 20
gallon.
In addition to the reservoir size, some factor should be considered when selecting
the reservoir; it must allow the air to escape and dirt to settle. Moreover, it must be able
to capture all the drained oil and the oil returned from the system. In addition, the oil
level must be high enough to avoid the entrained air getting into the system, which affects
the overall performance, and cause wear to the pumps. Lastly, it should have a large
surface area to allow system heat dissipation.
After discussing these considerations, the governing equations for pumps, motor,
and the load become as following.
For each load and its counterpart pump, the governing equation becomes
Q pump  Qleakage  Qload
(3-31)
Q pump  Qleakage   volD
(3-32)
1
Qload  Qo  K c P
2
P
(3-33)
1 
1 
 volD  Qo 
Kc 
2 
(3-34)
45
After calculating coefficient for each pump, table 3-1 shows the coefficients
summary
Table 3- 1 Pumps and loads coefficients
Gear Pump 2,000
Gear Pump
Piston Pump
psi
1,500 psi
10,000 psi
0.85
0.85
1010,00010,000
0.95
1010,00010,000
0.8
0.8
Psi
0.9
Psi
3000
1750
1750
rpm
0.401
0.142
0.024
in3/rev
Kc
1.2133e-3
3.429e-4
1.5e-5
Qo
4
0.8
0.089
gpm
psi
gpm
Volumetric
efficiency
Total
efficiency
nominal
speed
volumetric
displacement
units
3-5 Controller
After steady state system analysis, system dynamics and controller should be
introduced. Controller is a set of components used to bound system dynamics within
certain limits depending on the operation requirements; these components type can be
either mechanical, electrical, hydraulic, or a combination of them.
Before designing the controller, the complete model and the operation procedure
should be introduced. Figure 3-8 and 3-9 show the system model and the circuit operation
flow chart.
46
Figure 3- 7 Overall system simulation model
47
Figure 3- 8 Circuit operation flow chart
The circuit starts with checking the main control valve status (power take off
valve). In case the valve is open, the controller check both pumps loading status; if one or
both of pump are not loaded, the controller open the unloading valve associated with the
unloaded pump to minimize the consumed power. Once the pump is loaded, the
controller estimates the required pump speed based on the pressure difference,
consequently two required speed associated each pump are generated. Because both of
48
pumps speeds are related to each other with the gear reduction, only one signal has to be
adjusted at a time. The controller selects the signal from the pump that has higher
deviation from the required pressure. Based on the actual motor speed, the controller
produces a control signal to the control valve solenoid to change the motor swash plate
angle, and accordingly changing the motor shaft speed. The sequence is repeated until
one of the following cases is presented: the main control valve is closed, the highpressure accumulator pressure is low, or the accumulator available pressure cannot
operate the circuit with the motor at maximum volumetric displacement. Once the circuit
is stopped, the controller closes the main control valve and opens the unloading valves.
The controller consists of two smaller controllers, the first is to produce the
desired speed based on the pressure difference and the other is to produce the input
current to the control valve based on the input speed. The first one cannot work properly
unless the second one is already tuned, as the swash plate angle reflects the speed output,
hence the pressure. Based on this introduction, the second controller will be designed
first.
For the first controller, figure 3-10 shows the construction of the motor controller
block diagram. The desired speed ωd is the input to the system and the difference ε is fed
to the controller that generates input current i based on the input. This current is fed to the
control valve’s solenoid that produces an equivalent swash angle α to control the motor
speed ω.
49
Figure 3- 9 Motor controller block diagram
The characteristic equation of the system is
1  Gs  0
(3-35)
Where Gs 
C1C2 K e S  Ti C1C2
S 3  S 2
S 3  S 2  C1C2 K e S  Ti C1C2  0 as   1
(3-36)
The equation becomes of second order where
 n 2  Ti C1C 2 , 2n  C1C2 K e
(3-37)
By selecting appropriate values for  , n ; n  25rad sec and   1
The natural frequency selected value is low because high frequencies can create
irritating noise problems and premature failures of vibrating parts. The damping ratio is
selected of a value 1 to ensure that there is no overshoot; as high values of either currents
or pressure can cause damage to the operating systems if they pass the allowed values
[11].
50
After calculations Ti  3e  3 and K e  2.5e  4. Figure 3-11 shows the motor circuit
block diagram.
Figure 3- 10 Motor circuit block diagram
The second step is to design the controller for generating the speed signal based
on the pressure difference. This part of the controller will consist of two smaller
controllers for each pump.
For the previous circuit, a block reduction method should be conducted [5];
figures 3-12 and 3-13 shows the overview of the circuit including the controller for the
low pressure pump and ENERPAC pump respectively.
Figure 3- 11 Low pressure pump control circuit
51
Figure 3- 12 ENERPAC pump control circuit
PID controller design has several methods like pole placement, root locus, and
optimization using linear quadratic [4], however PID controller design is difficult in
design and it is used when PI controller cannot operate the circuit properly in term of
stability and response. PI controller was tested first then stability was checked for the
closed loop function using MATLAB SISO tool and if PI controller caused stability
problems, PID controller would be used.
The transfer function of the system with the PI controller is
P  ke S  ki

Pd  S

C1 S 2  C 2



3
2
 C3 S  S  C 4 S  C5 
 ke

 S 1


ki
C1 S 2  C 2
P




or
 ki
 S  C3 S 3  S 2  C 4 S  C5 
Pd




(3-38)
The closed loop transfer function becomes
C3 S 4  1  k e C2 S 3  ( K i C1  C4 ) S 2  (C5  K e C2 ) S  K i C2
(3-39)
Where C1= 58.819, C2=705.834, C3=3e-3, C4=48.366, C5=580.39 for the low pressure
pump
C1= 73.69, C2=884.323, C3=3e-3, C4=48.366, C5=580.39 for the first stage pump
52
C1= 318.252, C2=3819.016, C3=3e-3, C4=48.366, C5=580.39 for the high pressure pump
After using SISO tool, PI controller showed stability and met the system requirements.
Table 3-2 shows the PI coefficient for each pump to work properly.
Table 3- 2 Pumps controller coefficients
Gear Pump 2,000 psi
Piston Pump 10,000 psi
Gear Pump 1,500 psi
Ke
0.636
0.084375
1
Ki
3.18
0.54
4.55
After system analysis, steady state analysis, system dynamics, and controller
design the cycle time has to be determined. The cycle time was determined by the load
requirements. The single acting cylinder and the pole chainsaw represented the load; both
of them operated for 60 seconds. For the single acting cylinder, it operated in the first
stage for 29 sec and in the second stage 31 sec. For the pole chainsaw, there were two
cases to consider, the first case at which the pole chainsaw operated for one minute and
the second case at which it operated for 20 seconds every one minute.
53
Chapter 4 SIMULATION MODEL
In this chapter, the model for hydrostatic transmission driving the accessories tool
using power take-off in MATLAB/SIMULINK version R2008b is presented.
Figure 4-1 shows the complete simulation model overview, the model has five
blocks: controller, control valve, hydraulic motor, 2,000 Psi pump, and 10,000 Psi pump
or ENERPAC pump.
Figure 4- 1 Overall system simulation model
Starting from the loads and their counterpart pumps figures 4-2, 4-3, 4-4, and 4-5
shows the low pressure pump, ENEPRAC PUMP, quick extend pump, and high-pressure
pump models with their counterpart load respectively.
54
Figure 4- 2 Low-pressure pump 2,000 psi simulation model
Figure 4- 3 ENERPAC pump simulation model
Figure 4- 4 Quick extend stage pump 1,500 Psi simulation model
55
Figure 4- 5 High-pressure pump 10,000 pump simulation model
As the pumps shafts are driven by hydraulic motor, figure 4-6 shows the hydraulic
motor model.
Figure 4- 6 Hydraulic motor simulation model
The accumulator model represented the power source, in form of pressure, from
the vehicle side; the motor volumetric displacement determines the flow rate. Based on
the accumulator available pressure and the motor volumetric displacement, that is
determined by the swash plate angle, the motor output toque changes.
The hydraulic accumulator output pressure changes as its fluid volume changes,
figure 4-7 shows the hydraulic accumulator model.
56
Figure 4- 7 Hydraulic accumulator simulation model
As the motor drives the pumps, the load torque changes depending on the pump
output pressure, accordingly changes the motor output speed. Figure 4-8 shows the load
torque model.
Figure 4- 8 Load torque simulation model
The motor swash plate angle changes depending on the control valve output.
Figure 4-9 shows the control valve model.
57
Figure 4- 9 Control valve simulation model
As shown, the control valve output depends on the input current. The input
current depends on the controller signal. Figure 4-10 shows the controller model.
Figure 4- 10 Controller simulation model
The controller has 3 blocks; low pressure controller, ENERPAC pressure
controller, and RPM controller. The controller measures the pressure difference from
both pumps and based on that difference it generates the required output speed in RPM.
The speed is compared with the actual speed of the controller and based on the difference
58
it generates the required current to be fed to the control valve. In addition, the controller
generates control signal to control the unloading valves and the power take off valve
based on the loads signals. Figure 4-11, 4-12, and 4-13 show the low-pressure pump,
ENERPAC pump, and RPM controller model respectively.
Figure 4- 11 Low-pressure pump 2,000 psi controller simulation model
Figure 4- 12 ENERPAC pump 1,500/10,000 psi controller simulation model
59
Figure 4- 13 RPM controller simulation model
This chapter discussed in details the complete model, details of each component,
and the operation procedure.
60
Chapter 5 SIMULATION RESULTS
In this chapter, simulation results are presented and discussed. The results
included pressure change for each pump, accumulator status in term of pressure and flow
rate, Total power drawn during the cycle time, control valve solenoid input current, and
swash plate angle for the two cases. The first case is when the pole chainsaw is operated
for the whole operation cycle time-one minute, the quick extend stage operates for 29 sec,
and the high-pressure stage for 31 sec and it is referred to as case 1. The second case at
which the pole chainsaw is operated for 20 seconds every one minute for quick extend
and high-pressure stages period and it is referred to as case 2. The discussion included
model results verification with the manufacturer’s datasheet. Pressure was expressed in
psi, flow rate in gallon per minute (gpm), torque in N.m, power in horsepower (hp), and
current in Ampere (amp).
5-1 Simulation results for case 1
For the pumps pressure output with the motor speed, figure 5-1 shows the output
pressures for case 1.
61
Figure 5- 1 Output pressure VS. motor speed for case 1
The high-pressure pump has two stages, quick extend at low pressure and highpressure stage. At the end of the quick extend period and the beginning of high-pressure
application, because of small change in the load torque 0.15 N.m, the motor experiences
small change in the speed. The controller changes the swash plate angle to match new
load toque based on the accumulator available pressure. Figures 5-2 and 5-3 show the
change in the input current and the swash plate angle for case 1 and the change in the
torque respectively.
62
Figure 5- 2 Control valve input current and output swash plate angle case 1
63
Figure 5- 3 Motor output torque and pumps input torque case 1
As the high-pressure accumulator is the main hydraulic power source for the
accessories power system, the power required within the driving cycle has to be within
the accumulator capabilities. Figure 5-3 and 5-4 show the status of the accumulator
within the driving cycle with the motor input power and the final status of the
accumulator respectively.
64
Figure 5- 4 Accumulator outputs case 1
Figure 5- 5 Accumulator final state case 1
The total fluid drawn from the accumulator in case 1 is 3.57 gallon and the final
pressure of the accumulator is 3055 psi while the total power used is 490.3 hp in one
minute.
5-2 Simulation results for case 2
For the pumps pressure output with the motor speed, figure 5-6 shows the output
pressures for case 2.
65
Figure 5- 6 Output pressure VS. motor speed for case 2
As shown, at the end of the pole chainsaw supply period, the pump output
pressure is directed to the tank and become zero gauge pressure. Because of changing the
load torque significantly, the motor experiences a change in the speed before changing
the swash plate angle to match new load toque using the available pressure; figures 5-7
and 5-8 show the change in the input current and the swash plate angle for case 2 and the
change in the torque respectively.
66
Figure 5- 7 Control valve input current and output swash plate angle case 2
Figure 5- 8 Motor output torque and pumps input torque case 2
67
From the previous figures, by the end of the low-pressure pump period, the
decrease in load torque making the controller to decease the required swash plate angle
with decreasing the input current.
As presented, after 20 seconds there is a spike in motor speed; this is because of
the great change in the load torque that is much less than the motor torque and this
difference is converted into speed. In the high pressure stage, the decrease in speed is
because the required torque is more than provided by the motor and this causes an
instantaneous decrease in motor output speed.
The next Figures 5-9 and 5-10 show the status of the accumulator within the
driving cycle with the motor input power and the final status of the accumulator
respectively.
68
Figure 5- 9 Accumulator outputs case 2
Figure 5- 10 Accumulator final state case 2
The total fluid drawn from the accumulator in case 2 is less than the first case
1.295 gallon and the final pressure of the accumulator is 4109 psi while the total power
used is 206.5 hp in one minute.
To test accumulator capabilities to drive the system with the available hydraulic
motor, other cases were conducted. The results for the these cases were as following:
69

The accumulator can operate the low pressure pump with 2,000 psi alone for 159
seconds with input power consumption rate 7.085 hp/sec and overall efficiency
72.64%

The accumulator can operate the ENERPAC pump alone with 29 seconds for
quick extend and the rest as high pressure stage for 1322 seconds with input
power consumption rate 1.058 hp/sec and overall efficiency 80.98%.

The accumulator can operate the ENERPAC pump alone in quick extend stage
only for 1128 seconds with input power consumption rate 1.241 hp/sec and
overall efficiency 72.65%.

The accumulator can operate the ENERPAC pump alone in high pressure stage
only for 1333 seconds with input power consumption rate 1.05 hp/sec and overall
efficiency 81.19%.

The accumulator can operate case 1 for 120.3 seconds with input power
consumption rate 8.168 hp/sec and overall efficiency 73.48%.

The accumulator can operate case 2 for 240.9 seconds with input power
consumption rate 3.447 hp/sec and overall efficiency 74.94%.
The following equation represents the available accumulator capabilities to operate
the circuit with the available hydraulic motor
1.241* t quick _ extend  1.05 * t high_ pressure  7.085 * tlow_ pressure  1399 hp
t quick _ extend  1128
Where
t high_ pressure  1327
tlow_ pressure  159
70
(5-1)
All time values are in seconds.
In addition to the previous loading cases, More loading cases are included in
Appendix A with different loading pressures.
Model verification is based on three criteria: analyze the model performance
based on what was designed for in term of final speed and pressure, test the component
assumed efficiencies reflection on the output, and compare the input power consumption
rate to the values on the manufacturer’s datasheet.
From the previous figures, the system reaches the required pressure for different
cases. Also the system reflects the accumulator status in different time ranges and loads.
The results overall efficienies reflected the assumed efficiency for both pumps
and hydraulic motor. For example, for low pressure pump as the overall efficiency was
80% and the motor overall effeciency was 90%, the efficiency has to be around 72 %.
The efficency reported was 72.64 %.
For the input power consumption rate, from the datasheet the low pressure pump
at pressure 2000 psi, and input speed 3000 RPM, the required input was 7.5 hp which is
equal to (7.5/0.8=9.375) hp after excluding the pump total efficiency. The low pressure
pump consumed power reported was 9.8402 with error 4.96% from the datasheet value.
For ENEPRAC pump, the error in quick extend stage was 10.22%; for the high pressure
stage, the error was 5.5%. This error was due to propagation of small error in converting
pressure, torque, and speed units. Also, due to the components assumed efficiecnies, this
propagation error increased at lower efficiencies. The error was higher in case of
ENERPAC pump than the low pressure pump, with the same assumed efficiencies,
71
becuase the first pump operated for 1128 seconds while the other worked for 159 seconds
only.
5-3 Summary
In this chapter, outputs from each pump to its counterpart load, accumulator
power used, and controller outputs were presented. Case 1 represented the output at
which the system at full load with both pumps in operation for one minute. Case 2
represented the output at which the pole chainsaw pump work in a portion of the cycle
time. Table 5-1 shows the comparison between two cases in term of power used, final
pressure in the accumulator, and the total fluid drawn.
Table 5- 1 Simulation summary
Case 1
Case 2
units
Accumulator final pressure
3055
4109
Psi
Total fluid used
3.57
1.295
gpm
Total power used
490.3
206.5
hp
72
Chapter 6 SUMMARY AND FUTURE WORK
6-1 Summary
This research provided a comprehensive study for hydraulic systems that can be
used to drive the hydraulic accessories from the power take off for HHV. Three hydraulic
driving systems were proposed: hydraulic intensifier, hydraulic transformer, and
hydrostatic transmission system. Preliminary analysis was conducted on the proposed
systems to select the most efficient operating system for the hydraulic tools with the least
impact on the main power source.
The hydraulic intensifier new design satisfied the isolation test, however it
required intensive maintenance to ensure complete isolation. Moreover, the efficiency in
the working condition was low, 55%; therefore, it was not suitable for this application.
Hydraulic transformer has high efficiency. In addition, the isolation could be
attained by installing a separate driving system not with using the power source directly
from the power take off point and this is not the scope of this research.
The hydrostatic transmission has the least impact on the HHV as it isolated the
HHV power source from the PTO system. The motor converted the hydraulic energy into
mechanical energy to drive pumps with separate tanks. In addition, using axial piston
73
variable displacement motor improved the system efficiency. Accordingly, the
hydrostatic
system
was
analyzed
and
designed
then
simulated
using
MATLAB/SIMULINK to operate the hydraulic load properly.
The analysis started with defining the pressure working ranges, then the design of
the pumps, and hydraulic motor with the control valve. The system model included
hydraulic accumulator as the main power source from the HHV into the model.
The model included components data from manufacturer’s datasheets so that the
model outputs represented real application output values.
It was operated at the
maximum operating point and the components efficiencies are the values at this point.
The result included several cases in order to define the system capabilities. For
each case, it provided the total power consumption rate and the maximum operation time
cycle for the system using the available main circuit power.
The model was verified by comparing the consumed power to the manufacturer’s
datasheet values and comparing the assumed efficiencies to the model output efficiencies.
6-2 Conclusion
The model presented the operation of the hydrostatic transmissions system to
drive the hydraulic accessories using HHV power through power take-off.
Swash plate piston pumps should be used over gear pumps because of high
efficiency. However, these pumps are expensive comparing to gear pumps. Therefore, the
economic perspective of the proposed system versus their performance should also be
evaluated via experimental research.
74
The reduction gear relates both pumps speeds which decreases the degree of
freedom. It must be selected so that both designed operating points require the same
motor shaft speed.
The operated pumps should be close in volumetric displacement so that unloading
one of them does not break the other pump’s shaft due to the instantaneous high motor
torque. This effect will increase with on-off cycles.
6-3 Future work
This research represents the foundation for the accessories system for the HHV as
it provides a comprehensive study of driving systems for the hydraulic accessories.
Experimental tests should be conducted on the hydrostatic transmission in order to verify
the proposed mathematical model. With experiment, the complete analysis of the pumps’
and
hydraulic
motor’s
efficiencies
will
75
increase
the
model
accuracy.
References
[1] Filipi, K. Y. (2004). Simulatoin study of a series hydraulic hybrid propulsion system for a light
truck.
[2] Merritt, H. E. (1991). Hydraulic Control Systems.
[3] Stojkov, B. T. (1997). The Valve Primer.
[4] Astrom, K., & Hagglund, T. (1995). PID controller: Theory, Design, and Tuning.
[5] Ogata, K. (2009). Modern Control Engineering 5th Edition.
[6] MARE, P. J. (n.d.). SIMPLIFIED MODEL OF PRESSURE REGULATED, VARIABLE DISPLACEMENT
PUMPS FOR THE SIZING OF COMPLEX HYDRAULIC SYSTEMS.
[7] Pourmovahed, A., Beachley, N., & Fromczak, F. (1992). Modeling of a hydraulic Energy
regeneration system.
[8] Mathworks SimHydraulics. SimHydraulics user guide.
[9] Wikipedia. (n.d.). Retrieved from http://en.wikipedia.org/wiki/Bulk_modulus.
[10] Parr, A. (1998). Hydraulic and pnuematics- A technician's and Engineer guide- 2nd Edition.
[11] Manring, N. D. (2005). Hydraulic system control.
[12] Manring, N. (1996). Modeling and Desigining a variable displacement open loop pump.
[13] Shan, M. (2009). Modeling and Control Strategy for Series Hydraulic Hybrid Vehicles.
[14] Manring, N. (1997). The Effective Fluid Bulk-Modulus Within a Hydrostatic Transmission.
[15] Cheng, C. (2010). Application of Artificial Neural Networks in the Power Split Controller for a
Series Hydraulic Hybrid Vehicle.
[16] STANELY hydraulic tools. (n.d.).
http://www.stanleyhydraulic.com/Products/CircleSawsCR/tabid/99/Default.aspx.
76
[17] Manring, N. (1998). Modeling and Designing a Hydrostatic Transmission With a FixedDisplacement Motor.
[18] Rexroth Bosch Group-Hydraulic servos. (n.d.). Retrieved from
http://www.boschrexroth.com/country_units/america/united_states/sub_websites/brus_brh_i
/en/products_ss/07_proportional_servo_valves/06_servo_valves/index.jsp.
[19] Huskey Tools. (n.d.). Retrieved from http://www.huskietools.com/Catalog2009.pdf.
[20] Enerpac cylinders. (n.d.). Retrieved from http://www.enerpac.com/enUS/products/cylinders-lifting-products-and-systems/general-purpose/rc-series-single-actinggeneral-pur-0.
[21] Maxtor Jet. (n.d.). Retrieved from http://www.maximator-jet.de/.
[22] Hydraulics and Pneumatics. (n.d.). Retrieved from
http://www.hydraulicspneumatics.com/200/FPE/Pumps/Article/True/6404/Pumps.
[23] Carrlane. (n.d.). Retrieved from
http://www.carrlane.com/Catalog/index.cfm/27425071F0B221118070C1C512A0A1F0900101B0
30010543C1C0C16190D172D252A5E435E5D51.
[24] High Pressure Equipments HiP. (n.d.). Retrieved from
http://www.highpressure.com/pumping.asp?ID=5&ptype=hi&section=10.
[25] Enerpac Intensifier. (n.d.). Retrieved from http://www.enerpac.com/files/PID_E214US.pdf.
[26] Engineering ToolBox. (n.d.). Retrieved from http://www.engineeringtoolbox.com/bulkmodulus-elasticity-d_585.html?v=1.5e3&units=Pa.
[27] Eaton. (n.d.). Retrieved from http://hydraulics.eaton.com/products/pdfs/E-PUGE-MC001E1.pdf.
[28] Innas Hydraulic Transformer. (n.d.). Retrieved from http://www.innas.com/IHT.html.
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[33] Abuhaiba, M. (2009). Mathematical Modeling and Analysis of a Variable Displacement
Hydraulic Bent Axis Pump Linked to High Pressure and Low Pressure Accumulators.
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77
[35] Hydraulic and Pnuematics. (n.d.).
http://www.hydraulicspneumatics.com/200/TechZone/Accumulators/Article/True/6446/TechZo
ne-Accumulators.
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http://www.enerpac.com/files/catalogues/ZE_options325US.pdf.
78
Appendix A
Table A- 1 Maximum operating time for different cases
Load Pressure (psi)
Kc (psi/ gpm)
Maximum Operating
time( seconds)
1000
0.00242
395.2
1200
0.00201667
329.3
1400
0.00172857
274
1600
0.0015125
227.8
1800
0.00134444
190.4
7500
1.7097E-05
1504
8000
1.6028E-05
1410
Eaton Gear Pump only
ENERPAC Piston Pump
only
79
Figure A- 1 Loading pressure Vs. maximum operating time for Eaton pump
80
Appendix B
Table B- 1 Hydraulic tools input power requirement
Equipment
Company Pressure (psi)
Abrasive cut off saw
StanelyHydraulic
1500-2000
7-9 gpm
Abrasive cut off saw
StanelyHydraulic
n/a
10-15 gpm
Abrasive cut off saw
StanelyHydraulic
n/a
7-9 gpm
Abrasive cut off saw
StanelyHydraulic
n/a
7-9 gpm
Underwater Grinder
StanelyHydraulic
n/a
4-10 gpm
Grinder
StanelyHydraulic
n/a
7-9 gpm
Cupstone Grinder
StanelyHydraulic
n/a
10 gpm
Bull-Nose Grinder
StanelyHydraulic
n/a
5-10 gpm
Horizontal Grinder
StanelyHydraulic
n/a
8-10 gpm
Horizontal Grinder
StanelyHydraulic
n/a
8-10 gpm
n/a
5-10 gpm
Cut-off Saw
Crowder
Hydraulic
81
Flow
Tools
Cut-off Saw
Crowder
Hydraulic
Tools
n/a
5-8 gpm
Hydraulic Drills
StanelyHydraulic
n/a
5.8-13 gpm
Hydraulic Drills
StanelyHydraulic
n/a
12 gpm
Hydraulic Drills
StanelyHydraulic
n/a
3-9 gpm
Hydraulic Drills
StanelyHydraulic
n/a
7-9 gpm
Hydraulic Drills
StanelyHydraulic
n/a
7-9 gpm
Hydraulic Drills
StanelyHydraulic
n/a
7-9 gpm
Hydraulic Drills
StanelyHydraulic
n/a
7-9 gpm
Hydraulic Rock Drill
Crowder
Hydraulic
Tools
1450-2000 psi
5.3-6.6 gpm
Core Drill
Crowder
Hydraulic
Tools
1160-2500 psi
5.3 gpm
Core Drill
Crowder
Hydraulic
Tools
1160-2500 psi
5.3 gpm
Rock Drill
Crowder
Hydraulic
Tools
1450-2000 psi
5.3-6.6 gpm
Pick Hammer
Crowder
Hydraulic
Tools
1000-1300 psi
5.3 gpm
Rod Driver
Crowder
Hydraulic
Tools
2000 psi
5-8 gpm
82
Rod Driver
Crowder
Hydraulic
Tools
2000 psi
5-8 gpm
Hydraulic Breaker
StanelyHydraulic
1500-2000
7-9 gpm
Hydraulic Breaker
Sunrise
1280-1850
3-7
Hydraulic Breaker
Sunrise
1140-1700
5-8
Hydraulic Chipper
StanelyHydraulic
1000-2000
8 gpm
Hydraulic Chipper
StanelyHydraulic
1000-2000
8 gpm
Hydraulic Chipper
StanelyHydraulic
1000-2000
5 gpm
Hydraulic Chipper
StanelyHydraulic
1000-2000
5 gpm
Hydraulic Chipper
StanelyHydraulic
1000-2000
4-6 gpm
Hydraulic Digger
StanelyHydraulic
1000-2000
7-9 gpm
Diamond Chain Saw
StanelyHydraulic
n/a
4-6 gpm
Diamond Chain Saw
StanelyHydraulic
n/a
7-9 gpm
Diamond Chain Saw
StanelyHydraulic
n/a
7-9 gpm
Diamond Chain Saw
StanelyHydraulic
n/a
12
Earth Auger
StanelyHydraulic
n/a
7-9 gpm
Earth Auger
StanelyHydraulic
n/a
7-9 gpm
Ground Rod Driver
StanelyHydraulic
n/a
7-9 gpm
n/a
7-9 gpm
Ground Rod Driver
Stanely-
83
Hydraulic
Ground Dod Driver
StanelyHydraulic
n/a
5-9 gpm
Ground Dod Driver
StanelyHydraulic
n/a
5-9 gpm
Post Puller
StanelyHydraulic
n/a
7-9 gpm
Post Puller
StanelyHydraulic
n/a
7-9 gpm
Post Puller
StanelyHydraulic
n/a
7-9 gpm
Post Puller
StanelyHydraulic
n/a
3-9 gpm
Post Driver
Crowder
Hydraulic
Tools
400
5-8 gpm
Post Driver
Crowder
Hydraulic
Tools
400
5-8 gpm
Post Driver
Crowder
Hydraulic
Tools
200
5-8 gpm
Post Driver
Crowder
Hydraulic
Tools
150
5-8 gpm
Impact wrenches
StanelyHydraulic
n/a
4-12 gpm
Impact wrenches
StanelyHydraulic
n/a
4-12 gpm
Impact wrenches
StanelyHydraulic
n/a
4-12 gpm
Impact wrenches
StanelyHydraulic
n/a
4-12 gpm
Impact wrenches
StanelyHydraulic
N/A
4-12 gpm
84
Impact wrenches
StanelyHydraulic
N/A
Impact wrenches
StanelyHydraulic
N/A
Impact wrenches
StanelyHydraulic
N/A
Impact wrenches
StanelyHydraulic
N/A
Impact wrenches
StanelyHydraulic
N/A
Chainsaw
Sunrise
1000-2000
5-8
Chainsaw
Sunrise
1000-2000
5-8
Chainsaw
Sunrise
1000-2000
4-6
Chainsaw
Sunrise
1000-2000
4-8
Chainsaw
Sunrise
1000-2000
4-8
Chainsaw
Sunrise
1000-2000
4-6
Chain Saw
Stanley
Hydraulic
Tools
1000-2000
7-9 gpm
Circular Saw for tree
trimming
Stanley
Hydraulic
Tools
1000-2000
5-7
Pole Chainsaw
Stanley
Hydraulic
Tools
1000-2000
4-6
Pole Chainsaw
Stanley
Hydraulic
Tools
1000-2000
4-6
Pole Chainsaw
Stanley
Hydraulic
Tools
1000-2000
4-6
Pole Chainsaw
Stanley
Hydraulic
Tools
1000-2000
7-9
4-12 gpm
7-12 gpm
7-12 gpm
7-12 gpm
7-12 gpm
85
Pole Chainsaw
Stanley
Hydraulic
Tools
1000-2000
7-9
Pole Chainsaw
Stanley
Hydraulic
Tools
1000-2000
7-9
Pruner
Stanley
Hydraulic
Tools
1000-2000
3-9
Pruner
Stanley
Hydraulic
Tools
1000-2000
3-9
Extende Prunner
(Scissor Style)
Stanley
Hydraulic
Tools
1000-2000
N/A
Abrasive cut off saw
StanelyHydraulic
1100-2150
5.3-10.6
Post Hole Borer
Crowder
Hydraulic
Tools
1300-1600
5.3
Paving BreaKers
Crowder
Hydraulic
Tools
1300-1600
5.3
Paving BreaKers
Crowder
Hydraulic
Tools
1300-1600
5.3
Paving BreaKers
Crowder
Hydraulic
Tools
1500-1800
7.9
Paving BreaKers
Crowder
Hydraulic
Tools
1500-1800
7.9
Paving BreaKers
Crowder
Hydraulic
Tools
1500-1800
7.9
Paving BreaKers
Crowder
Hydraulic
Tools
1500-1800
7.9
86
Paving BreaKers
Crowder
Hydraulic
Tools
1500-1800
7.9
Paving BreaKers
Crowder
Hydraulic
Tools
1300-1600
5.3
Paving BreaKers
Crowder
Hydraulic
Tools
1300-1600
5.3
Paving BreaKers
Crowder
Hydraulic
Tools
1300-1600
5.3
Paving BreaKers
Crowder
Hydraulic
Tools
1500-1800
7.9
Paving BreaKers
Crowder
Hydraulic
Tools
1500-1800
7.9
Paving BreaKers
Crowder
Hydraulic
Tools
1500-1800
7.9
Paving BreaKers
Crowder
Hydraulic
Tools
1500-1800
7.9
Paving BreaKers
Crowder
Hydraulic
Tools
1500-1800
7.9
Paving BreaKers
Crowder
Hydraulic
Tools
1500-1800
7.9
Paving BreaKers
Crowder
Hydraulic
Tools
1500-1800
7.9
Paving BreaKers
Crowder
Hydraulic
Tools
N/A
31(max)-610(regulated)
Oil Divider
Crowder
Hydraulic
up to 2000
7-9 gpm
87
Tools
Auger (Post Hole
Digger)
Stanley
Hydraulic
Tools
up to 2000
7-9 gpm
Self-Contained
Hydraulic Cutters
Enerpac
Max. 10,000
N/A
Self-Contained
Hydraulic Cutters
Enerpac
Max. 10,000
N/A
Self-Contained
Hydraulic Cutters
Enerpac
Max. 10,000
N/A
Self-Contained
Hydraulic Cutters
Enerpac
Max. 10,000
N/A
Self-Contained
Hydraulic Cutters
Enerpac
Max. 10,000
N/A
Self-Contained
Hydraulic Cutters
Enerpac
Max. 10,000
N/A
Self-Contained
Hydraulic Cutters
Enerpac
Max. 10,000
N/A
Jack, Hydraulic Toe
Enerpac
Max. 10,000
N/A
Jack, Hydraulic Toe
Enerpac
Max. 10,000
N/A
Hydraulic Machine
Lifts
Enerpac
Max. 10,000
N/A
Hydraulic Machine
Lifts
Enerpac
Max. 10,000
N/A
PMU Series Torque
Wrench Pump
Enerpac
700(1st stage)-10,000(2nd stage)
200 in3/min(1st
stage)-20
in3/min(2nd stage)
PMU Series Torque
Wrench Pump
Enerpac
700(1st stage)-10,000(2nd stage)
200 in3/min(1st
stage)-20
in3/min(2nd stage)
PMU Series Torque
Wrench Pump
Enerpac
700(1st stage)-11,600(2nd stage)
200 in3/min(1st
stage)-20
in3/min(2nd stage)
PMU Series Torque
Wrench Pump
Enerpac
700(1st stage)-11,600(2nd stage)
88
200 in3/min(1st
stage)-20
in3/min(2nd stage)
S Series Square
Drive Torque
Wrenchs
Enerpac
max. 10,000 psi
N/A
S Series Square
Drive Torque
Wrenchs
Enerpac
max. 10,000 psi
N/A
S Series Square
Drive Torque
Wrenchs
Enerpac
max. 10,000 psi
N/A
S Series Square
Drive Torque
Wrenchs
Enerpac
max. 10,000 psi
N/A
S Series Square
Drive Torque
Wrenchs
Enerpac
max. 10,000 psi
N/A
Flow Control Valves
Enerpac
rated for 10,000 psi
N/A
Flow Control Valves
Enerpac
rated for 10,000 psi
N/A
Flow Control Valves
Enerpac
rated for 10,000 psi
N/A
Flow Control Valves
Enerpac
rated for 10,000 psi
N/A
Flow Control Valves
Enerpac
rated for 10,000 psi
N/A
Flow Control Valves
Enerpac
rated for 10,000 psi
N/A
Flow Control Valves
Enerpac
rated for 10,000 psi
N/A
Flow Control Valves
Enerpac
rated for 10,000 psi
N/A
3-Way Directional
Control Valves
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
gal/min
3-Way Directional
Control Valves
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
gal/min
3-Way Directional
Control Valves
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
gal/min
3-Way Directional
Control Valves
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
gal/min
3-Way Directional
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
89
Control Valves
gal/min
3-Way Directional
Control Valves
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
gal/min
3-Way Directional
Control Valves
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
gal/min
3-Way Directional
Control Valves
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
gal/min
4-Way Directional
Control Valves
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
gal/min
4-Way Directional
Control Valves
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
gal/min
4-Way Directional
Control Valves
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
gal/min
4-Way Directional
Control Valves
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
gal/min
4-Way Directional
Control Valves
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
gal/min
4-Way Directional
Control Valves
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
gal/min
4-Way Directional
Control Valves
Enerpac
max. 10,000 psi
Flow Capcaity 4.5
gal/min
Grease Control
Valves Heavy Duty
High Pressure
Lincoln
max 7500 psi
n/a
Grease Control
Valves Heavy Duty
High Pressure
Lincoln
max 7500 psi
n/a
6 Ton Remote
Compression CHead w/Rubber
Boot( page 52 )
6.2 Ton Dieless
Remote Crimping
Head with Rubber
Boot – .9” Jaw
Opening (page 53)
n/a
Huskie tools
10,000
n/a
Huskie tools
10,000
90
6.2 Ton Flip-Top
Dieless Remote
Crimping Head –
1.5” Jaw Opening
12 Ton Remote
Compression CHead with Rubber
Boot – 1” Jaw
Opening
12 Ton Remote
Titanium
Compression CHead – 1” Jaw
Opening ( page 54 )
12 Ton Remote
Compression CHead with Rubber
Boot – 1-3/16” Jaw
Opening (page 55)
12 Ton Remote
Compression CHead with Rubber
Boot – 1.65” Jaw
Opening ( page 55)
12 Ton Remote
Compression CHead with Rubber
Boot – 1.5” Jaw
Opening ( page 56)
n/a
Huskie tools
10,000
n/a
Huskie tools
10,000
n/a
Huskie tools
10,000
n/a
Huskie tools
10,000
n/a
Huskie tools
10,000
n/a
Huskie tools
10,000
15 Ton Remote
Compression CHead – 2” Jaw
Opening (page 56)
n/a
Huskie tools
10,000
60 Ton Compression
Head
Huskie tools
10,000
60 Ton Compression
Head
Huskie tools
10,000
60 Ton Double
Acting Compression
Head
Huskie tools
10,000
n/a
n/a
n/a
91
60 Ton Double
Acting Compression
Head
Huskie tools
10,000
100 Ton
Compression Heads
Huskie tools
10,000
200 Ton
Compression Heads
Huskie tools
10,000
Hydraulic Scissor
Cutter and Remote
Head ( page 74)
Huskie tools
10,000
Hydraulic Scissor
Cutter and Remote
Head ( page 75)
Huskie tools
10,000
Hydraulic Scissor
Cutter and Remote
Head ( page 76)
Huskie tools
10,000
Hydraulic Scissor
Cutter and Remote
Head ( page 76)
Huskie tools
10,000
Hydraulic Low
Pressure 4” Cable
Cutter
Huskie tools
2,500
35 Ton Hydraulic
Steel Punch
Huskie tools
10,000
6 Ton Remote
Compression CHead w/Rubber
Boot( page 52 )
n/a
n/a
n/a
n/a
n/a
n/a
n/a
n/a
n/a
n/a
Huskie tools
10,000
6.2 Ton Dieless
Remote Crimping
Head with Rubber
Boot – .9” Jaw
Opening (page 53)
Huskie tools
10,000
SH-70A Hydraulic
Power Puncher
Kudos
700 bar
82 CC
Kudos
700 bar
87.4 cc
Hydraulic Flange
Spreaders & Pipe
n/a
92
Bender
Bus Bar Cutter
Kudos
700 bar
116 cc
Bus Bar Bender
Kudos
700 bar
124 cc
Hydraulic Flange
Puller
FastorQ
10,000
Hydraulic Flange
Puller
FastorQ
10,000
Chipping Hammer
Fairmont
1,400 - 1,700 PSI
5 GPM
Breaker
Fairmont
1,280 - 1,850 PSI
3 - 7 GPM
Breaker
Fairmont
1,140 - 1,700 PSI
5 - 8 GPM
Breaker
Fairmont
5 - 8 GPM
Breaker
Fairmont
5 - 8 GPM
Breaker
Fairmont
7 - 9 GPM
Pistol-Grip Chain
Saw
Fairmont
1,000 - 2,000 PSI
4 - 8 GPM
Standard Chain Saw
Fairmont
1,000 - 2,000 PSI
4 - 8 GPM
Long Reach Chain
Saw
Fairmont
1,000 - 2,000 PSI
5 - 8 GPM
Crimping Tool
Fairmont
1,400 - 2,500 PSI
3 - 9 GPM
Crimping Tool
Fairmont
1,500 - 2,500 PSI
3 - 9 GPM
Crimping Tool
Fairmont
1,500 - 2,500 PSI
3 - 9 GPM
Dieless Crimping
Tool
Fairmont
1,400 - 2,500 PSI
3 - 9 GPM
Crimping Tool
Fairmont
1,500 - 2,500 PSI
3 - 9 GPM
Crimping Tool
Fairmont
1,500 - 2,500 PSI
3 - 9 GPM
Crimping Tool
Fairmont
10,000 PSI
Cut-Off Saw
Fairmont
1,000 - 2,000 PSI
5 - 8 GPM
Overhead Circular
Saw
Fairmont
1,000 - 2,000 PSI
4 - 6 GPM
93
n/a
n/a
Drill
Fairmont
1,000 - 2,000 PSI
4 - 8 GPM
Hammer Drill
Fairmont
1,000 - 2,000 PSI
5 - 10 GPM
Rock Drill
Fairmont
1,275 - 1,700 PSI
5 - 7 GPM
Ground Rod Driver
Fairmont
2,000 PSI Max
5 - 8 GPM
Impact Wrench
Fairmont
1,000 - 2,500 PSI
4 - 10 GPM
Impact Wrench
Fairmont
1,000 - 2,500 PSI
4 - 10 GPM
Impact Wrench
Fairmont
1,000 - 2,500 PSI
4 - 10 GPM
Impact Wrench
Fairmont
1,000 - 2,500 PSI
4 - 10 GPM
Impact Wrench
Fairmont
1,000 - 2,500 PSI
4 - 10 GPM
Impact Wrench
Fairmont
1,000 - 2,500 PSI
4 - 10 GPM
Sign Post Puller
Fairmont
1,000 - 2,000 PSI
4 - 6 GPM
Pole Puller
Fairmont
300 - 2,800 PSI
4 - 15 GPM
Pole Tamper
Fairmont
1,000 - 2,000 PSI
4 - 6 GPM
Pole Tamper
Fairmont
1,000 - 2,000 PSI
4 - 6 GPM
Utility Pruner
Fairmont
1,000 - 2,000 PSI
4 - 6 GPM
Orchard and Shade
Tree Pruner
Fairmont
1,000 - 2,000 PSI
4 - 6 GPM
2" Submersible
Pump
Fairmont
0 - 2,000 PSI
5 - 8 GPM
2.5" Submersible
Pump
Fairmont
0 - 2,000 PSI
4 - 8 GPM
3" Submersible
Pump
Fairmont
up to 2,500 PSI
7 - 10 GPM
3" Submersible
Trash Pump
Fairmont
up to 2,500 PSI
7 - 10 GPM
4" Submersible
Pump
Fairmont
0 - 2,000 PSI
5 - 9 GPM
Rotamag Rail Drill
Racine
Railroad
Products
2,000 PSI
10 GPM
94
Rail Profile Grinder
Racine
Railroad
Products
2,000 PSI
10 GPM
Bull Nose Grinder
Racine
Railroad
Products
2,000 PSI
5 GPM
Rail Wled Shear
Racine
Railroad
Products
2,000 PSI
5 - GPM
Saw Slotter
Racine
Railroad
Products
2,000 PSI
5 GPM
Rail Puller
Racine
Railroad
Products
2,000 PSI
10 GPM
Back Handle Grinder
Racine
Railroad
Products
2,000 PSI
5 - 10 GPM
Tie Tamper
Racine
Railroad
Products
2,000 PSI
10 GPM
Spike Driver
Racine
Railroad
Products
2,000 PSI
10 GPM
Spike Puller
Racine
Railroad
Products
2,000 PSI
5 - 10 GPM
Impact Drill
Racine
Railroad
Products
1,000 - 2,000 PSI
5 GPM
Right Angle Grinder
Racine
Railroad
Products
2,000 PSI
5 GPM
Cup Stone Grinder
Racine
Railroad
Products
2,000 PSI
10 GPM
Sprint Saw
Racine
Railroad
2,000 PSI
10 GPM
95
Products
Stand-Up Web
Grinder
Racine
Railroad
Products
2,000 PSI
10 GPM
1" Impact Wrench
Racine
Railroad
Products
2,000 PSI
10 GPM
Concrete Chain Saw
Reimann &
Georger
Corp
2,000 - 2,500 PSI
8 GPM
Circular Saw
(Hydrasaw)
Reimann &
Georger
Corp
2,000 - 2,500 PSI
5 - 8 GPM
Core Drill
Reimann &
Georger
Corp
2,000 - 2,500 PSI
8 - 15 GPM
Hydra Breaker
Reimann &
Georger
Corp
2,000 PSI
5 - GPM
Post Driver
Reimann &
Georger
Corp
2,000 PSI
5 - 8 GPM
Hydra Pump
Reimann &
Georger
Corp
2,000 - 2,500 PSI
5 - 8 GPM
Post Puller
Reimann &
Georger
Corp
2,500 PSI
2 - 10 GPM
Hand Chain Saw
Reimann &
Georger
Corp
1,000 - 2,000 PSI
4 - 8 GPM
Pole Saw
Reimann &
Georger
Corp
1,000 - 2,000 PSI
5 - 8 and 4 - 6
GPM
Impact Wrench
Reimann &
Georger
Corp
1,000 - 2,500 PSI
4 - 12 GPM
96
Atlas Copco LH 11
Hydraulic Pick
Hammer
Ohio Power
Tool
1,000 - 1,300 PSI
5.3 GPM
Atlas Copco LH 18
Lightweight Breaker
Ohio Power
Tool
1,300 - 1,600 PSI
5.3 GPM
Atlas Copco LH 22
Medium Breaker
Ohio Power
Tool
1,500 - 1,800 PSI
5 - 8 GPM
Atlas Copco LH 27
Heavy-duty Breaker
Ohio Power
Tool
1,500 - 1,800 PSI
5 - 8 GPM
Atlas Copco LH 39
Super Heavy-duty
Breaker
Ohio Power
Tool
1,500 - 1,800 PSI
8 - 10 GPM
Atlas Copco LHD 23
M Hydraulic Rock
Drill
Ohio Power
Tool
1,450 - 2,000 PSI
5.3 - 6.6 GPM
Atlas Copco LCD
1500 Hydraulic Core
Drill
Ohio Power
Tool
2,200 PSI
5.8 GPM
Atlas Copco LCD
500 Hydraulic Core
Drill
Ohio Power
Tool
2,200 PSI
5.8 GPM
Atlas Copco LS 14
Hydraulic Cut-Off
Saw
Ohio Power
Tool
2,500 PSI
5 - 8 GPM
Atlas Copco LS 16
Hydraulic Cut-Off
Saw
Ohio Power
Tool
2,500 PSI
5 - 10 GPM
Atlas Copco LTP 3
Hydraulic Water
Pump
Ohio Power
Tool
2,000 - 4,200 PSI
6.9 - 10.1 GPM
Atlas Copco LWP 2
Hydraulic Water
Pump
Ohio Power
Tool
1,500 - 4,000 PSI
4.8 - 6.4 GPM
97